HEAT   ENGINES 

STEAM,  GAS,  STEAM  TURBINES 
AND  THEIR  AUXILIARIES 


BY 
JOHN  R.  ALLEN 

PROFESSOR   MECHANICAL   ENGINEERING 
UNIVERSITY   OF   MICHIGAN 

-AND- 

JOSEPH  A.  BURSLEY 

JUNIOR   PROFESSOR   MECHANICAL   ENGINEERING 
UNIVERSITY   OF   MICHIGAN 


SECOND  EDITION 
THOROUGHLY  REVISED  AND  ENTIRELY  RESET 


TOTAL  ISSUE  6,000 


McGRAW-HILL  BOOK  COMPANY,  INC, 
239  WEST  39TH  STREET,  NEW  YORK 

6  BOUVERIE  STREET,  LONDON,  E.  C. 
1914 


COPYRIGHT  1914,  BY  THE 
McGRAW-HiLL  BOOK  COMPANY,  INC. 


COPYRIGHT  1910,  BY  THE 
McGRAW-HiLL  BOOK  COMPANY. 


THE. MAPLE. PBESS. YORK. PA 


PREFACE  TO  SECOND  EDITION 

The  advancement  in  heat  engines  during  the  past  four  years 
has  been  such  as  to  make  it  desirable  to  publish  a  new  edition 
of  this  text.  Among  the  new  subjects  treated  are  the  Stumpf 
Uniflow  Engine,  the  Humphrey  Gas  Pump,  and  recent  develop- 
ments in  steam  turbines  and  gas  engines.  It  has  been  found  also 
desirable  to  rewrite  many  of  the  chapters  in  order  to  clear  up 
the  points  which  experience  has  shown  need  more  detailed  explana- 
tion than  was  given  in  the  first  edition.  At  the  same  time  an 
effort  has  been  made  not  to  increase  materially  the  size  of  the 
book. 

The  authors  desire  to  thank  those  members  of  the  faculty  of 
the  mechanical  engineering  department  of  the  University  of 
Michigan  who  have  assisted  in  the  preparation  of  this  text  by 
their  timely  suggestions,  and  the  manufacturers  who  have  kindly 
supplied  many  of  the  new  cuts  used. 

JOHN  R.  ALLEN. 
JOSEPH  A.  BURSLEY. 
ANN  ARBOR,  MICHIGAN 
Sept.  1,  1914. 


PREFACE  TO  FIRST  EDITION 

In  preparing  this  book,  it  has  been  the  intention  of  the  authors 
to  present  an  elementary  treatise  upon  the  subject  of  Heat 
Engines,  considering  only  those  engines  which  are  most  com- 
monly used  in  practice.  It  is  written  primarily  as  a  text-book, 
the  subject-matter  having  been  used  in  the  classes  at  the 
University  of  Michigan  for  a  number  of  years. 

The  forms  of  heat  engines  discussed  include  the  steam  engine 
with  its  boiler  plant  and  auxiliaries,  the  gas  engine  with  its  pro- 
ducer, oil  engines,  and  the  principal  types  of  steam  turbines. 
Under  each  division  of  the  text,  problems  have  been  worked  out 
in  detail  to  show  the  application  of  the  subject-matter  just 
treated,  and,  in  addition,  a  large  number  of  problems  have  been 
introduced  for  class-room  work.  The  use  of  calculus  and  higher 
mathematics  has  been  largely  avoided,  the  only  place  where  it 
is  used  being  in  the  chapter  on  thermo-dynamics,  which  subject 
has  been  treated  in  its  elementary  phases  only.  The  matter 
of  the  design  of  engines  has  been  left  untouched,  as  it  was  felt 
that  that  subject  did  not  properly  come  within  the  scope  of  this 
work. 

The  authors  wish  to  express  their  thanks  to  Messrs.  H.  C. 
Anderson,  A.  H,  Knight,  and  J.  A.  Mover,  for  their  assistance  in 
compiling  this  work,  to  Mr.  W.  R.  McKinnon  who  made  a  number 
of  the  drawings,  and  to  the  various  manufacturers  who  have  very 
kindly  furnished  illustrations  and  descriptions  of  their  apparatus. 

JOHN  R.  ALLEN. 
JOSEPH  A.  BURSLEY. 

ANN  ARBOR,  MICHIGAN 
Sept.  1,  1910. 


VI 


CONTENTS 

PAGE 

PREFACE y-v 

LIST  OF  TABLES xii 

CHAPTER  I 
HEAT 

THEORY  AND  MEASUREMENT  OF  HEAT 1 

SPECIFIC  HEAT 5 

RADIATION,  CONDUCTION,  AND  CONVECTION 7 

DEFINITIONS  OF  ENERGY,  WORK,  AND  POWER     .            8 

CHAPTER  II 
ELEMENTARY  THERMODYNAMICS 

FIRST  AND  SECOND  LAWS  OF  THERMODYNAMICS 10 

EQUATION  AND  LAWS  OF  PERFECT  GASES 11 

ABSORPTION  OF  HEAT 13 

JOULE'S  LAW 14 

RELATION  OF  SPECIFIC  HEATS .  15 

EXPANSIONS  IN  GENERAL 17 

WORK  OF  EXPANSION 18 

GENERAL  CASE  OF  HEAT  ADDED 20 

HEAT  ADDED  AT  CONSTANT  VOLUME  OR  CONSTANT  PRESSURE  22 

ADIABATIC  EXPANSION 24 

ISOTHERMAL  EXPANSION 25 

RELATION   OF   PRESSURE,  VOLUME   AND   TEMPERATURE  DURING  EX- 
PANSION   27 

THEORETICAL  HEAT  ENGINE " 29 

CARNOT  CYCLE 29 

PROBLEMS 34 

CHAPTER  III 
PROPERTIES  OF  STEAM 

FORMATION  OF  STEAM ^ 

PROPERTIES  OF  STEAM    . 40 

STEAM  TABLES     .      .  ^ 42 

CHAPTER  IV 
CALORIMETERS  AND  MECHANICAL  MIXTURES 

CALORIMETERS     

SEPARATING  CALORIMETERS 

THROTTLING  CALORIMETERS 50 

vii 


viii  CONTENTS 

PAGE 

QUALITY  OF  STEAM    .            53 

PROBLEMS 55 

MECHANICAL  MIXTURES 56 

PROBLEMS 61 

CHAPTER  V 
COMBUSTION  AND  FUELS 

COAL  ANALYSIS 64 

HEATING  VALUE  OF  FUELS 66 

COAL  CALORIMETERS 68 

AIR  REQUIRED  FOR  COMBUSTION 70 

SMOKE 71 

ANALYSIS  OF  FLUE  GASES 71 

THEORETICAL  TEMPERATURE  OF  COMBUSTION 74 

FUELS 75 

PROBLEMS 79 

CHAPTER  VI 
BOILERS 

RETURN  TUBULAR  BOILERS 82 

INTERNALLY  FIRED  BOILERS 87 

WATER-TUBE  BOILERS 91 

HORSE-POWER  OF  BOILERS 97 

HEATING  SURFACE,  GRATE  SURFACE,  AND  BREECHING 98 

BOILER  ECONOMY 99 

BOILER  EFFICIENCY 100 

BOILER  LOSSES ! 101 

BOILER  ACCESSORIES 105 

CHAPTER  VII 
BOILER  AUXILIARIES 

MECHANICAL  STOKERS 109 

INCLINED  GRATE  STOKERS 110 

CHAIN  GRATE  STOKERS 113 

UNDER-FEED  STOKERS 115 

BOILER  FEEDERS '.......  119 

FEED-WATER  HEATERS 122 

ECONOMIZERS 125 

SUPERHEATERS     127 

CHIMNEYS 128 

MECHANICAL  DRAFT 132 

PROBLEMS                                                                             133 


CONTENTS  ix 

CHAPTER  VIII 
STEAM  ENGINES 

PAGE 

THE  SIMPLE  STEAM  ENGINE 139 

HORSE-POWER  OF  A  STEAM  ENGINE 142 

LOSSES  IN  THE  STEAM  ENGINE 144 

METHODS  OF  REDUCING  INITIAL  CONDENSATION 147 

CLEARANCE  AND  COMPRESSION 149 

PROBLEMS 150 

CHAPTER  IX 
TYPES  AND  DETAILS  OF  STEAM  ENGINES 

CLASSIFICATION  OF  ENGINES 153 

ENGINE  DETAILS 160 

CHAPTER  X 
TESTING  OF  STEAM  ENGINES 

STEAM  ENGINE  INDICATOR 168 

INDICATED  HORSE-POWER 172 

DETERMINATION  OF  INITIAL  CONDENSATION        174 

ECONOMY  OF  VARIOUS  FORMS  OF  ENGINES 177 

BRAKE  HORSE-POWER 178 

MECHANICAL  EFFICIENCY 179 

ACTUAL  HEAT  EFFICIENCY 180 

DUTY 180 

PROBLEMS 184 

CHAPTER  XI 
VALVE  GEARS 

PLAIN  D-SLIDE  VALVES 187 

DEFINITIONS  OF  LAP,  LEAD,  ANGULAR  ADVANCE,  AND  ECCENTRICITY     .  188 

RELATIVE  POSITION  OF  VALVE  AND  PISTON 190 

ZEUNER  VALVE  DIAGRAM 191 

EFFECT  OF  CONNECTING  ROD 196 

VARIOUS  TYPES  OF  VALVES 198 

CORLISS  VALVES 205 

REVERSING  GEARS 209 

VALVE  SETTING 212 

INDICATOR  DIAGRAMS 213 

CHAPTER  XII 
GOVERNORS 

TYPES 216 

GOVERNOR  MECHANISM  .                 217 


x  CONTENTS 

PAGE 

ISOCHRONISM 222 

HUNTING  . 222 

FLY  WHEEL 223 

CHAPTER  XIII 
COMPOUND  ENGINES 

COMPOUND  ENGINES 225 

TANDEM  COMPOUND  ENGINES 226 

CROSS  COMPOUND  ENGINES 227 

RATIO  OF  CYLINDERS .      .      ...      .      .      .  228 

HORSE-POWER  OF  COMPOUND  ENGINES 229 

COMBINED  INDICATOR  CARDS 232 

PROBLEMS 234 

CHAPTER  XIV 
CONDENSERS  AND  AIR  PUMPS 

JET  CONDENSERS .      .      ...      .   •  .      .      .      .  237 

SURFACE  CONDENSERS 240 

AIR  PUMPS .      . 240 

COOLING  WATER 240 

PROBLEMS 242 

CHAPTER  XV 

.  STEAM  TURBINES 

HISTORY 244 

CLASSIFICATION 246 

ACTION  OF  STEAM  IN  TURBINE 247 

TURBINE  NOZZLES 248 

SPEED  OF  TURBINES 249 

DE  LAVAL  TURBINE 251 

CURTIS  TURBINE 255 

RATEAU  TURBINE 257 

KERR  TURBINE 259 

STURTEVANT  TURBINE 260 

PARSONS  TURBINE 262 

DOUBLE-FLOW  TURBINE 264 

LOW-PRESSURE  TURBINES 266 

MIXED-FLOW  TURBINES 267 

CHAPTER  XVI 
GAS  ENGINES 

HISTORY 269 

CLASSIFICATION 270 

THEORETICAL  EFFICIENCY 275 

LOSSES  282 


CONTENTS  xi 

PAGE 

GAS-ENGINE  FUELS 

GAS  PRODUCERS  . 
LIQUID  FUELS 
FUEL  MIXTURES  . 
RATED  HORSE-POWER      . 
ACTUAL  HORSE-POWER   . 

CHAPTER  XVII 
DETAILS  OF  GAS-ENGINE  CONSTRUCTION 

DESCRIPTION  OF  PARTS  . 
METHODS  OF  IGNITION  . 
METHODS  OF  GOVERNING  .... 

CARBURETORS       

OIL  ENGINES  FOR  SHIPS 
HUMPHREY' GAS  PUMP 
PROBLEMS        .... 

CHAPTER  XVIII 
ECONOMY  OF  HEAt  ENGINES 

RELATIVE  ECONOMY  .      .      . 

COMMERCIAL  ECONOMY 31° 

INDEX.  313 


LIST  OF  TABLES 

PACK 

TABLE  I.  TEMPERATURE  COLORS 4 

TABLE  II.  SPECIFIC  HEATS  OF  GASES 6 

TABLE  III.  RADIATING  POWER  OF  BODIES 7 

TABLE  IV.  CONDUCTING  POWER  OF  BODIES 8 

TABLE  V.  HEAT  AND  TEMPERATURE  CHANGES  DEPENDENT 

UPON  VALUE  OF  n  DURING  EXPANSION 18 

TABLE  VI.  SPECIFIC  HEATS  OF  SUPERHEATED  STEAM 40 

TABLE  VII.  PROPERTIES  OF  SATURATED  STEAM 42 

TABLE  VIII.  SPECIFIC  HEATS  OF  LIQUIDS  AND  SOLIDS 57 

TABLE  IX.  COMBUSTION  PROPERTIES  OF  ELEMENTS 67 

TABLE  X.  CALORIFIC  VALUE  OF  WOODS 76 

TABLE  XI.  CALORIFIC  VALUE  OF  PEATS 77 

TABLE  XII.  CALORIFIC  VALUE  OF  LIGNITES 77 

TABLE  XIII.  CALORIFIC  VALUE  OF  BITUMINOUS  COALS 78 

TABLE  XIV.  CALORIFIC  VALUE  OF  SEMI-BITUMINOUS  COALS  ...  78 

TABLE  XV.  CALORIFIC  VALUE  OF  ANTHRACITE  COALS 79 

TABLE  XVI.  DIAMETER  OF  BOILER  TUBES - .  99 

TABLE  XVII.  HEAT  BALANCE  IN  BOILER  PLANT 101 

TABLE  XVIII.  CHIMNEY  HEIGHTS 131 

TABLE  XIX.  STEAM  CONSUMPTION  OF  VARIOUS  CLASSES  OF  ENGINES  178 

TABLE  XX.  DUTY  OF  VARIOUS  FORMS  OF  PUMPS 181 

TABLE  XXI.  RELATIVE  CHANGES  IN  VELOCITY,  SPECIFIC  VOLUME 

AND  PRESSURE  OF  STEAM  FLOWING  THROUGH  A 

NOZZLE 249 

TABLE  XXII.  CALORIFIC  VALUE  OF  GASEOUS  AND  LIQUID  FUELS.  .  287 

TABLE  XXIII.  VOLUMETRIC  EFFICIENCIES,  vvj  OF  GAS  ENGINES  .  .  289 

TABLE  XXIV.  ECONOMIC  EFFICIENCIES,  vw,  AND  AIR  CONSUMPTION, 

OF  GAS  ENGINES 290 

TABLE  XXV.  THERMAL  EFFICIENCIES  OF  PRIME  MOVERS  ....  309 

TABLE  XXVI.  COMPARATIVE  COSTS  PER  RATED  HORSE-POWER  .  311 


xni 


HEAT  ENGINES 

STEAM-GAS-STEAM    TURBINES-AND 
THEIR  AUXILIARIES 

CHAPTER  I 
HEAT 

1.  Heat  being  the  source  of  energy  for  the  devices  considered  in 
this  book,  a  short  discussion  of  the  nature  and  the  more  important 
properties  of  heat  will  assist  the  student  to  a  better  understanding 
of  the  subject-matter  of  this  text.     These  phenomena  will  be  con- 
sidered only  as  they  affect  perfect  gases,  steam,  and  water. 

2.  Theory  of  Heat. — The  accepted  theory  of  heat  at  the  pres- 
ent time  is  that  it  is  a  motion  of  the  molecules  of  a  body.     Phys- 
ical experiments  indicate  this  to  be  the  fact.     The  intensity  of 
the  heat,  or  the  temperature,  is  supposed  to  depend  upon  the 
velocity  and  amplitude  of  these  vibrations. 

Most  bodies  when  heated  expand.  This  expansion  is  probably 
due  to  the  increased  velocity  of  the  molecules  which  forces  them 
farther  apart  and  increases  the  actual  size  of  the  body. 

The  vibration  may  become  so  violent  that  the  attraction  be- 
tween the  molecules  is  partly  overcome  and  the  body  can  no 
longer  retain  its  form.  In  this  case  the  solid  becomes  a  liquid. 
If  still  more  heat  is  added,  the  attraction  of  the  molecules  may 
be  entirely  overcome  by  their  violent  motion,  and  the  liquid 
then  becomes  a  gas. 

The  phenomena  of  heat  is  then  a  form  of  motion.  This  is  often 
stated  in  another  way,  that  is,  heat  is  a  form  of  kinetic  energy. 
As  heat  is  a  form  of  motion,  it  must  be  possible  to  transform  heat 
into  mechanical  motion.  In  the  following  pages,  therefore,  the 
most  important  methods  of  making  this  transformation  will  be 
discussed. 

3.  Temperature  and  Temperature  Measurement. — The   ve- 
locity of  the  vibration  of  the  molecules  of  a  body  determines 
the  intensity  of  the  heat,  and  this  intensity  is  measured  by 

1 


2       A :  f •'*: ;  • :      ;  i  ff&AT  ENGINES 

the  temperature.  If  the  molecules  of  a  body  move  slowly  it 
is  at  a  low  tempertaure;  if  they  move  rapidly  it  is  at  a  high  tem- 
perature. The  temperature  of  a  body  is  then  determined  by 
the  rapidity  of  the  motion  of  its  molecules. 

Temperature  is  sometimes  defined  as  the  thermal  state  of  a 
body  considered  with  reference  to  its  ability  to  transmit  heat  to 
other  bodies.  Two  bodies  are  said  to  be  at  the  same  temperature 
when  there  is  no  transmission  of  heat  between  them.  If  there  is 
transmission  of  heat  between  them,  the  one  from  which  the  heat 
is  flowing  is  said  to  have  the  higher  temperature. 

In  mechanical  engineering  work,  temperatures  are  usually 
measured  on  the  Fahrenheit  scale,  and  in  this  text,  unless 
otherwise  stated,  the  temperature  will  be  taken  on  this 
scale.  There  is,  however,  an  increasing  use  of  the  Centigrade 
scale  among  engineers,  and  certain  quantities,  such  as  the  in- 
crease in  temperature  in  a  dynamo,  are  always  expressed  in 
Centigrade  units. 

In  the  Fahrenheit  scale  the  graduations  are  obtained  by  noting 
the  position  of  the  mercury  column  when  the  bulb  of  the  ther- 
mometer is  placed  in  melting  ice,  and  again  when  it  is  placed  in 
boiling  water  under  an  atmospheric  pressure  corresponding  to 
sea  level  barometer.  The  distance  between  these  two  points  is 
divided  into  180  equal  parts.  The  freezing  point  is  taken  as 
32°,  making  the  boiling  point  32°  -f  180°  =  212°  above  zero. 

In  the  Centigrade  scale  the  distance  between  the  freezing  point 
and  the  boiling  point  is  divided  into  100  equal  parts  or  degrees, 
and  the  freezing  point  on  the  scale  is  marked  0°.  The  boiling 
point  is  then  100°. 

Both  the  Fahrenheit  and  Centigrade  scales  assume  an  arbitrary 
point  for  the  zero  of  the  scale. 

Since  in  the  Fahrenheit  scale  there  are  180  divisions  between 
the  freezing  and  boiling  points  and  on  the  Centigrade  100  divi- 
sions, it  follows  that  1°  F.  =  f  °  C.,  or  1°  C.  =  f°  F.  As,  how- 
ever, the  freezing  point  on  the  Fahrenheit  scale  is  marked  32  and 
on  the  Centigrade  scale  0,  it  is  necessary  to  take  account  of  this 
difference  when  converting  from  one  scale  to  the  other.  If  the 
temperature  Fahrenheit  be  denoted  by  tF  and  the  temperature 
Centigrade  by  tc,  then  the  conversion  from  one  scale  to  the 
other  may  be  made  by  the  following  equations : — 

*F=**c  +  32;  (1) 

*c  =  f  (tF  -  32).  (2) 


HEAT  3 

The  measurement  of  temperature  is  not  so  simple  a  process 
as  is  generally  supposed.  The  mercury  of  the  ordinary  glass 
thermometer  does  not  expand  equal  amounts  for  equal  incre- 
ments of  heat,  and  the  bore  of  the  thermometer  is  not  abso- 
lutely uniform  throughout  the  whole  length  of  the  tube.  These 
inaccuracies  must  be  allowed  for  by  accurate  calibration.  In 
measuring  the  temperatures  of  liquids,  the  depth  to  which  the 
thermometer  is  immersed  affects  the  reading,  and  it  should  be 
calibrated  at  the  depth  at  which  it  is  to  be  used.  If  a  ther- 
mometer is  used  to  measure  the  temperature  of  the  air  in  a 
room  in  which  there  are  objects  at  a  higher  temperature,  its 
bulb  must  be  protected  from  the  radiant  heat  of  those  hot 
bodies.  When  accurate  temperature  measurements  are  desired, 
a  careful  study  should  be  made  of  the  errors  of  the  instrument 
and  the  errors  in  its  use. 

The  ordinary  form  of  mercury  thermometer  is  used  for  tem- 
peratures ranging  from  —40°  F.  to  500°  F.  For  measuring  tem- 
peratures below  — 40°  F,  thermometers  filled  with  alcohol  are  used. 
These  are,  however,  not  satisfactory  for  use  at  high  temperatures. 
When  a  mercury  thermometer  is  used  for  temperatures  above 
500°  F.,  the  space  above  the  mercury  is  filled  with  some  inert 
gas,  usually  nitrogen  or  carbon-dioxide,  placed  in  the  ther- 
mometer tube  under  pressure.  As  the  mercury  rises,  the  gas 
pressure  is  increased  and  the  temperature  of  the  boiling  point  of 
the  mercury  is  raised,  so  that  it  is  possible  to  use  these  ther- 
mometers for  temperatures  as  high  as  1000°  F.  This  is  the 
limit,  however,  as  the  melting  point  of  glass  is  comparatively 
low. 

For  temperatures  exceeding  800°  F.,  some  form  of  pyrometer  is 
generally  used.  The  simplest  of  these  is  the  metallic  or  mechan- 
ical pyrometer.  This  consists  of  two  metals  having  different  rates 
of  expansion,  such  as  iron  and  brass,  attached  to  each  other  at 
one  end  and  with  the  other  ends  free.  By  a  system  of  levers  and 
gears  the  expansion  of  the  metals  is  made  to  move  a  hand  over  a 
dial  graduated  in  degrees.  This  should  not  be  used  for  tempera- 
tures over  1500°  F. 

There  are  two  types  of  electrical  pyrometers  in  use  to-day.  In 
one,  the  thermo-electric  couple  is  employed  and  the  difference  in 
temperature  of  the  junctions  of  the  two  metals  forming  the  couple 
produces  an  electric  current  which  is  proportional  to  this  dif- 
ference, and  which  is  measured  on  a  galvanometer  calibrated  in 


4  HEAT  ENGINES 

degrees.     By  keeping  one  junction  at  a  known  temperature,  the 
other  may  be  computed.     This  may  be  used  up  to  2500°  F. 

The  second  type,  the  electrical  resistance  pyrometer,  depends 
upon  the  increase  in  electrical  resistance  of  metals  due  to  a  rise 
in  temperature. 

For  still  higher  temperatures  the  optical  pyrometer  gives  the 
most  satisfactory  results.  This  is  based  on  the  results  of  ex- 
periments made  by  Pouillet  which  show  that  incandescent 
bodies  have  for  each  temperature  a  definite  and  fixed  color,  as 
follows: — 

TABLE  I.  TEMPERATURE  COLORS 


Color 

Temp.  C. 

Temp.  F. 

Faint  red  

525 

977 

Dark  red  

700 

1292 

Faint  ( 

herry 

f-     «•  •   800         •: 

.    1472 

Cherry 

900 

1652 

Bright 

cherry  

1000 

1832 

Dark  orange  

1100 

2012 

Bright  orange  

1200 

2192 

White  heat  

1300 

2372 

Bright 

white  

1400 

2552 

Dazzling  white  

f  1500 
\  1600 

(2732 
\2912 

4.  Absolute  Zero. — In  considering  heat  from  a  theoretical 
standpoint,  it  is  necessary  to  have  some  absolute  standard  of 
comparison  for  the  scale  of  temperature,  so  that  the  absolute 
scale  is  largely  used. 

A  perfect  gas  contracts  TQI~~C  of  its  volume  at  32°  F.  for  each 

degree  that  it  is  reduced  in  temperature.  Hence  if  the  tempera- 
ture be  lowered  to  a  point  491.6°  below  32°,  its  volume  will  be- 
come zero.  This  point  is  called  the  absolute  zero  and  is  mani- 
festly an  imaginary  one.  (The  lowest  point  so  far  actually 
reached  by  experiment  is  about  —488.9°  F.)  For  ordinary  usage 
it  is  sufficiently  accurate  to  consider  absolute  zero  as  492°  below 
the  freezing  point  in  the  Fahrenheit  scale.  In  other  words,  to 
convert  to  the  absolute  scale,  add  460  to  the  temperature  ex- 
pressed in  degrees  Fahrenheit.  In  this  text  absolute  tempera- 
tures will  be  denoted  by  T  and  temperatures  in  degrees  Fahren- 
heit by  t. 

On  the  Centigrade  scale  the  absolute  zero  is  273.1°  below  the 


HEAT  5 

freezing  point,  and  for  all  practical  purposes,  temperatures  on 
the  absolute  scale  may  be  found  by  adding  273  to  the  thermometer 
reading  expressed  in  degrees  Centigrade. 

5.  Unit  of  Heat. — Heat  is  not  a  substance,  and  it  cannot  be 
measured  as  we  would  measure  water,  in  pounds  or  cubic  feet, 
but  it  must  be  measured  by  the  effect  which  it  produces.     The 
unit  of  heat  used  in  mechanical  engineering  is  the  heat  required 
to  raise  a  pound  of  water  one  degree  Fahrenheit.     The  heat 
necessary  to  raise  a  pound  of  water  one  degree  does  not  remain 
the   same   throughout   any   great  range   of  temperature.     For 
physical  measurements  where  accuracy  is  required,  it  is  neces- 
sary to  specify  at  what  point  in  the  scale  of  temperatures  this 
one  degree  is  to  be  taken.     The  practice  of  different  authors 
varies;  the  majority,  however,  specify  that  the  heat  unit  is  the 
amount  of  heat  required  to  raise  a  pound  of  water  from  39°  to  40° 
Fahrenheit.     The  range  from  39°  F.  to  40°  F.  is  used  because  at 
this  temperature  water  has  its  maximum  density.     This  unit  is 
called  a  British  Thermal  Unit,  and  is  denoted  by  B.T.U.     The 
heat  unit  used  in  Marks  and  Davis  tables  is  the  "mean  B.T.U.," 
that  is  j-J-o  of  the  heat  required  to  raise  one  pound  of  water  from 
32°  to   212°  at .  atmospheric  pressure  (14.7  pounds  per  square 
inch  absolute). 

6.  Specific  Heat. — If  the  temperature  of  a  body  is  raised  or 
lowered  a  definite  amount,  a  definite  amount  of  heat  must  either 
be  added  to  or  given  up  by  the  body. 

Then 

dH  =  Cdt.-  (3) 

where  C  is  the  heat  necessary  to  change  the  temperature  of  the 
body  one  degree. 

Let  a  body  of  unit  weight  at  a  temperature  T\  be  heated  to  a 
temperature  T2)  and  at  the  same  time  let  its  heat  content  be 
increase  from  Hi  to  Hz.  Then  the  heat,  H.  added  to  cause  this 
increase  in  temperature  will  be  found  by  integrating  equation 
(3)  between  the  limits  T\  and  T2,  or 


rTz 
H  =  H  2  —  HI  =    I  Cdt. 

JTi 


If  C  is  a  constant  and  is  equal  to  the  heat  necessary  to  raise 
the  temperature  of  a  unit  weight  one  degree 

rT2 
H  =  C  \  dt  =  C(T2  -  TO.  (4) 


6 


HEAT  ENGINES 


C  in  equation  (4)  represents  the  heat  capacity  of  the  body,  or 
the  heat  required  to  raise  the  temperature  of  a  unit  weight  of 
the  body  one  degree. 

The  heat  capacity  of  any  substance  compared  with  that  of 
an  equal  weight  of  water  is  called  its  specific  heat. 

Expressed  in  English  units,  the  heat  capacity  of  one  pound 
of  water  is  one  B.T.U.,  and  specific  heat  may  be  defined  as  the 
heat  necessary  to  raise  the  temperature  of  one  pound  of  a  substance 
one  degree  Fahrenheit  expressed  in  British  Thermal  Units. 

Since  a  B.T.U.  is  the  amount  of  heat  required  to  raise  a  pound 
of  water  from  39°  to  40°  the  specific  heat  will  then  represent  the 
ratio  of  the  heat  necessary  to  raise  the  temperature  of  a  unit 
weight  one  degree  to  the  heat  necessary  to  raise  the  temperature 
of  the  same  weight  of  water  from  39°  to  40°. 

TABLE  II.  SPECIFIC  HEATS  OF  GASES 


Gas' 

Symbol 

Expressed  in 
B.T.U 

Expressed  in 
Ft.  Lbs. 

« 

i 

A 
fc< 

05 

*>' 

X 

Constant 
pressure 

Cp 

Constant 
volume 
Cv 

Constant 
pressure 
KP 

Constant 
volume 
Kv 

Air  .    - 

.2375    .1689 
.453*   .400 
.5084    .350 
.2450    .174 
.2169;   .167 
.1569    .131 
.4797    .450 
3.40902.412 
.2438    .1727 
.2175    .1551 
j  VI  j  Page  40. 

184.77 
352.75 
395.54 
190.61 
168.75 
122.07 
373.21 
2652.20 
189.68 
169.22 

131.40 
311.20 

272  .  30 
135.37 
129.93 
101.92 
350.10 
1876.54 
134.36 
120.67 

53.37    .406 
41.55    .133 
123.24    .452 
55.24    .408 
38.82    .299 
20.15    .197 
23.11     .066 
775.66    .413 
55.32    .412 
48.55    .402 

Alcohol  

C2H6O 
NH3 
CO 
C02 

CS2 
C4HioO 
H 

! 

See  Tdbh 

Ammonia  gas  
Carbonic  oxide  
Carbonic  acid 

Carbon  disulphide.  . 
Ether  

Hydrogen  

Nitrogen  

OxvKen 

Superheated  steam. 

In  solid  and  liquid  substances  it  is  necessary  to  consider 
but  one  specific  heat,  as  the  change  in  volume  when  a  solid  or 
a  liquid  substance  is  heated  is  so  small  that  its  effect  may  be 
neglected.  In  gases  the  change  in  volume  when  the  gas  is 
heated  is  large,  and  if  it  is  heated  under  a  constant  pressure 
this  change  is  directly  proportional  to  the  change  in  the  abso- 
lute temperature.  If  there  is  a  change  in  volume  there  must 
be  external  work  done.  On  the  other  hand,  when  gas  is  con- 


HEAT  7 

fined  and  is  heated,  it  cannot  expand.  If  it  does  not  expand, 
there  is  no  external  work  done.  Therefore,  in  considering 
the  specific  heat  of  a  gas,  we  must  consider  two  cases:  one  in 
which  the  pressure  remains  constant  and  the  gas  expands  when 
it  is  heated;  and  the  other  where  the  volume  remains  constant 
and  the  pressure  increases  when  the  gas  is  heated.  Hence, 
in  the  case  of  a  gas,  there  are  two  specific  heats,  the  specific 
heat  of  constant  pressure  and  the  specific  heat  of  constant  volume. 
The  specific  heat  of  constant  volume  will  be  denoted  by  cv  and 
the  specific  heat  of  constant  pressure  by  cp,  both  being  expressed 
in  B.T.U.  When  expressed  in  foot-pounds  they  will  be  denoted 
by  Kv  and  Kp  respectively. 

7.  Radiation. — The  heat  that  passes  from  a  body  by  radia- 
tion may  be  considered  similar  to  the  light  that  is  radiated  from 
a  lamp.     There  is  always  a  transfer  of  radiant  heat  from  a  body 
of  a  high  temperature  to  a  body  of  lower  temperature.     The 
amount  of  heat  radiated  will  depend  upon  the  difference  in 
temperature  between  the  bodies  and  upon  the  substances  of 
which  they  are  composed.     The  following  table  gives  the  radi- 
ating power  of  different  bodies. 

TABLE  III.  RADIATING  POWER  OF  BODIES 

Radiating  power  of  bodies,  expressed  in  heat  units,  given  off  per  square  foot 
per  hour  for  a  difference  of  one  degree  Fahrenheit.     (PECLET.) 

B.T.U. 

Copper,  polished 0327 

Iron,  sheet 0920 

Glass 595 

Cast  iron,  rusted 648 

Building  stone,  plaster,  wood,  brick 7358 

Woolen  stuffs,  any  color 7522 

Water 1 .085 

8.  Conduction. — The  heat  transmitted  by  conduction  is  the 
heat   transmitted    through   the    body   itself.     The   amount   of 
heat  conducted  will  depend  upon  the  material  of  which  the 
body  is  composed  and  the  difference  in  temperature  between 
the  two  sides  of  the  body,  and  is  inversely  proportional  to  the 
thickness  of  the  body.     Heat  may  be  conducted  from  one  body 
to  another  when  they  are  placed  in  contact  with  each  other. 

The  following  table  gives  the  conducting  power  of  different 
bodies. 


8  HEAT  ENGINES 

TABLE  IV.  CONDUCTING  POWER  OP  BODIES 

The  conducting  power  of  materials,  expressed  in  the  quantity  of  heat 
units  transmitted  per  square  foot  per  hour  by  a  plate  one  inch  thick,  the 
surfaces  on  the  two  sides  of  the  plate  differing  in  temperature  by  one  degree. 
(PECLET.) 

B.T.U. 

Copper 515 

Iron 233 

Lead 113 

Stone 16.7 

Glass 6.6 

Brick  work 4.8 

Plaster 3.8 

Pine  wood .75 

Sheep's  wool 323 

9.  Convection. — Loss  by  convection  is  sometimes  called  loss 
by  contact  of  air.     When  air  or  other  gas  comes  in  contact 
with  a  hot  body  it  is  heated  and  rises,  carrying  away  heat  from 
the  body.     Heat  carried  off  in  this  manner  is  said  to  be  lost 
by  convection.     The  loss  by  convection  is  independent  of  the 
nature  of  the  surface — wood,  stone,   or  iron  losing  the  same 
amount — but  it  is  affected  by  the  form  and  position  of  the 
body. 

10.  Energy,  Work,  and  Power. — Work  is  the  overcoming  of 
resistance  through   space   and  is  measured   by  the  resistance 
multiplied  by  the  space  through  which  this  resistance  is  over- 
come.    The  simplest  form  of  work  is  the  raising  of  a  body 
against  the  force  of  gravity. 

Let  M    =  the  mass  of  the  body. 
g      =  the  force  of  gravity. 
w     =  the  weight. 

I       =  the  distance  through  which  the  weight  is  moved. 
W    =  work. 
Then  Mg  =  w,  and  wl  =  W. 

If  w  is  expressed  in  pounds  and  I  in  feet,  then  the  unit  of 
work  will  be  the  foot-pound  (ft.-lb.). 

If  we  consider  the  work  done  by  a  fluid,  let  the  volume  be 
increased  from  v  to  v  +  d  v,  and  the  pressure  against  which  the 
increase  takes  place  be  p,  then  the  work  done  will  be 

P  [(v  +  5  v)  -  v]  =  p  d  v  =  d  W. 


HEAT  9 

If  a  pressure  p  acts  upon  an  area  a  through  a  distance  I,  then 
the  work 

W  =  pla. 

Work  may  also  be  expressed  as  mass  times  acceleration  times 
space. 

Energy  is  the  capacity  for  doing  work. 

Power  is  the  time  rate  of  doing  work.  The  unit  of  power 
is  the  horse-power  (H.P.).  A  horse-power  is  equivalent  to  raising 
33,000  Ibs.  one  foot  in  one  minute.  This  is  the  unit  employed 
in  determining  the  power  of  a  steam  engine.  If  r  equals  the 
resistance  expressed  in  pounds,  I  the  distance  in  feet  through 
which  the  resistance  r  is  overcome,  and  m  the  time  in  minutes  in 
which  the  space  is  passed  over,  then  the  horse-power  exerted  is 

_J  Xr 
33,000  X  m' 

Power  is  often  expressed  in  electrical  units.  This  is  usually 
the  case  where  an  engine  is  used  to  drive  a  generator.  An 
ampere  is  the  unit  of  current  strength  or  rate  of  flow.  The 
volt  is  the  unit  of  electromotive  force  or  electrical  pressure. 
The  watt  is  the  product  of  the  amperes  and  the  volts.  One 
horse-power  equals  746  watts,  or  one  kilowatt  equals  1.34 
horse-powers. 


CHAPTER  II 
ELEMENTARY  THERMODYNAMICS 

11.  First    Law    of    Thermodynamics. — "'When    mechanical 
energy  is  produced  from  heat,  a  definite  quantity  of  heat  goes 
out  of  existence  for  every  unit  of  work  done;  and  conversely, 
when  heat  is  produced  by  the  expenditure  of  mechanical  energy, 
the  same  definite  quantity  of   heat  comes  into  existence  for 
every  unit  of  work  spent." 

The  relation  between  work  and  heat  was  first  accurately 
determined  by  Joule  in  1850.  More  recently  Professor  Rowland 
of  John  Hopkins  University  redetermined  its  equivalent  with 
great  accuracy.  His  results  show  that  one  British  Thermal  Unit 
is  equivalent  to  778  foot-pounds.  This  factor  is  often  called  the 
mechanical  equivalent  of  heat,  and  is  usually  denoted  by  J.  Heat 
and  work  are  mutually  convertible  in  the  ratio  of  778  foot-pounds 
equals  one  B.T.U. 

12.  Second  Law  of  Thermodynamics. — The  second  law  of 
thermodynamics   may  be   stated   in  different  ways.     Clausius 
states  it  as  follows:  "It  is  impossible  for  a  self-acting  machine, 
unaided  by  any  external  agency,  to  convey  heat  from  one  body 
to  another  of  higher  temperature.'7     Rankine  states  the  second 
law  as  follows:  "If  the  total  actual  heat  of  a  homogeneous  and 
uniformly   hot   substance  be   conceived  to  be   divided  into   a 
number  of  equal  parts,  the  effects  of  those  parts  in  causing  work 
to  be  performed  are  equal."     It  follows  from  the  second  law 
that  no  heat  engine  can  convert  more  than  a  small  fraction  of 
the  heat  given  to  it  into  work.     From  this  law  we  derive  the 
expression  for  the  efficiency  of  a  heat  engine,  i.  e., 

_  heat  added  —  heat  rejected 
heat  added 

The  second  law  is  not  capable  of  proof  but  is  axiomatic.  All 
our  experiments  with  heat  engines  go  to  show  that  this  law  is 
true. 

13.  Laws  of  Perfect  Gases. — There  are  two  laws  expressing 
the  relation  of  pressure,  volume,  and  temperature  in  a  perfect 
gas:  the  law  of  Boyle  and  the  law  of  Charles. 

10 


ELEMENTARY  THERMODYNAMICS  11 

Boyle's  Law.  —  "  The  volume  of  a  given  mass  of  gas  varies  in- 
versely as  the  pressure,  provided  the  temperature  remains  constant." 

If  p0  =  the  pressure,  and  v0  =  the  volume  of  the  initial 
condition  of  the  gas,  and  p  and  v  any  other  condition  of  the 
same  gas,  then 

POVO  =  pv  =  a  constant. 

Charles'  Law.  —  "  Under  constant  pressure  equal  volumes  of 
different  gases  increase  equally  for  the  same  increment  of  tem- 
perature. Also  if  the  gas  be  heated  under  constant  pressure 
equal  increments  of  its  volume  correspond  very  nearly  to  equal 
increments  of  temperature  by  the  scale  of  a  mercury  ther- 
mometer." 

This  law  may  also  be  stated  as  follows:  When  a  gas  receives 
heat  at  a  constant  volume  the  pressure  varies  directly  as  the  absolute 
temperature,  or  when  a  gas  receives  heat  at  a  constant  pressure  the 
volume  varies  directly  as  the  absolute  temperature. 

Letting  a  gas  receive  heat  at  a  constant  volume  v0,  the  pres- 
sure and  absolute  temperature  varying  from  p0,  T0  to  p,  T', 
then 


If  the  gas  now  receives  heat  at  this  pressure  p,  the  volume  and 
temperature  changing  to  v  and  T'  ',  then 

T 


14.  Equation  of  a  Perfect  Gas.  —  Combining  these  two  laws 
we  have  the  equation  of  a  perfect  gas.  Let  one  pound  of  a 
gas  have  a  volume  v,  a  pressure  p,  and  be  at  an  absolute  tem- 
perature T.  From  Boyle's  Law,  if  the  pressure  is  changed  to 
Pi  and  the  volume  to  v',  the  temperature  T  remaining  constant, 
then  we  have  the  following  equation: 


pi       v  pv 

-  =  -,'  or  Pl  = 

From  the  law  of  Charles,  if  the  volume  remains  constant  at 
vf  and  the  temperature  be  changed  to  Tr  and  the  pressure  to 
p'  ',  then 

Pl       T  p'T     «/ 

—f  =  ft'  or  PI  =  ~r' 


12  HEAT  ENGINES 

Combining  equations  (1)  and  (2),  we  have 

pv      p'T 

v'  ~~     Tf  ' 

Hence, 


p'v'     p"v"  f. 

=  ~rT  ••  =     v/  ==  a  constant  (3) 


Denoting  this  constant  by  R,  then 

pv  =  RT,  p'v'  =  RTr,  and  p"v"  =  RT".  (4) 

The  value  of  R  given  in  this  equation  is  for  one  pound  of  the 
gas.  If  we  wish  to  state  this  law  for  more  than  one  pound,  let 
w  equal  the  weight  of  the  gas,  then  the  law  becomes 

%  pv  =  wRT.  (5) 

This  equation  is  called  the  equation  of  the  gas  and  holds  true  for 
any  point  on  any  expansion  line  of  any  perfect  gas. 

These  laws  were  first  determined  for  air,  which  is  almost  a 
perfect  gas,  and  they  hold  true  for  all  perfect  gases.  A  perfect 
gas  is  sometimes  defined  as  a  gas  which  fulfils  the  laws  of  Boyle 
and  Charles.  It  is  probably  better  to  define  it  as  a  gas  in  which 
no  internal  work  is  done,  or  in  other  words,  a  gas  in  which 
there  is  no  friction  between  the  molecules  under  change  of 
conditions. 

In  the  above  expressions,  p  is  the  absolute  pressure  in  pounds 
per  square  foot,  0  is  the  volume  in  cubic  feet,  and  T  is  the  absolute 
temperature  in  degrees  Fahrenheit. 

Absolute  pressure  must  not  be  confused  with  gage  pressure. 
The  ordinary  pressure  gage  reads  the  difference  in  pressure  be- 
tween the  atmospheric  pressure  outside  the  gage  tube  and  the 
applied  pressure  inside  the  gage  tube.  The  absolute  pressure 
is  equal  to  the  gage  pressure  plus  the  barometric  pressure. 

The  value  of  R  for  any  given  substance  may  be  determined, 
provided  we  know  the  volume  of  one  pound  for  any  given  con- 
dition of  pressure  and  temperature.  For  example,  it  has  been 
found  by  experiment  that  for  air  under  a  pressure  of  14.7  Ibs. 
per  square  inch  absolute,  and  at  a  temperature  of  32°  F.,  the 
volume  of  1  Ib.  is  12.39  cu.  ft.  Substituting  these  values  in 
equation  (4)  we  have 


ELEMENTARY  THERMODYNAMICS  13 

R-V" 

K>    —     rp 

14.7  X  144  X  12.39 

32  +  460 

=  53.37    (compare   with    the   value   of  R  for   air   given  in 
Table  II).  (6) 

Therefore  for  one  pound  of  air  with  the  units  we  have  taken, 

pv  =  53.37  T.  (7) 

or  for  w  pounds, 

pv  =  53.37  wT.  (8) 

This  equation  is  always  true  for  air  at  all  times  and  under  all 
conditions,  as  long  as  it  remains  a  gas. 

Example. — A  tank*contains  5  Ibs.  of  air  at  75°  F.,  under  a  pressure  of 
100  Ibs.  per  square  inch  gage.     Find  the  volume  of  the  air. 
Solution. — 
pv  =  wRT. 

p  =  (100+  14.7)144  =  114.7  X  144  Ibs.  per  square  foot,  absolute. 
T  =  75  +  460  =  535°  absolute. 
Therefore,  substituting  in  the  equation  of  the  gas,  we  have 

114.7  X  144  X  v  =  5  X  53.37  X  535 

142760 
=    16520 

v  =  8 . 64  cu.  ft. 

Example.— Ten  pounds  of  air  under  a  pressure  of  50  Ibs.  per  square 
inch  gage  occupy  a  volume  of  10  cu.  ft.     Find  the  temperature. 
Solution. — 
pv  =  wRT 

p  =  (50  +  14.7)144  =  64.7  X  144  Ibs.  per  square  foot  absolute. 
Therefore 

64.7  X  144  X  10  =  10  X  53.37  X  T 

93200 
•  533.7 

T  =  174.5°  absolute 

T  =  174.5  -  460  =  -  285.5°  F. 

15.  Absorption  of  Heat. — When  a  gas  receives  heat  this  heat 
may  be  dissipated  in  one  or  all  of  three  ways;  by  increasing  its 
temperature,  by  doing  internal  work,  or  by  doing  external 
work. 


14  HEAT  ENGINES 

Let  dH  denote  the  heat  absorbed,  dS  the  heat  used  in  increas- 
ing the  temperature,  dl  the  heat  used  in  doing  internal  work, 
and  dW  the  heat  equivalent  of  the  external  work  done.  Then 

dH  =  dS  +  dl  +  dW  ^  (9) 

The  heat  utilized  in  changing  the  internal  energy  of  the  sub- 
stance is  represented  by  dS  -f  dl,  and  dl  -f-  dW  represents  the 
heat  equivalent  of  the  total  work  done. 

By  ''internal  work"  is  meant  work  done  in  overcoming  changes 
in  the  physical  state  of  the  substance,  and  in  overcoming  the 
attraction  of  the  molecules  for  each  other,  thus  changing  the  po- 
tential energy  of  the  body. 

An  example  of  this  is  shown  in  the  case  of  water  at  the  boiling 
point  being  changed  into  steam.  In  this  case  dS  in  equation 
(9)  becomes  zero  since  the  temperature  remains  constant.  There- 
fore all  the  heat  added  goes  to  doing  internal  and  external  work. 
The  external  work  will  be  equal  to  the  change  in  volume  from 
water  to  steam  times  the  pressure  under  which  the  steam  is  being 
formed.  This  will  be  only  a  small  part  of  the  total  heat  added  to 
accomplish  the  change,  the  balance  being  the  heat  going  to  in- 
ternal work  or  dl. 

Since  no  internal  work  is  done  in  heating  a  perfect  gas,  the  sec- 
ond term  in  equation  (9)  becomes  zero  and  all  the  heat  absorbed  * 
goes  either  to  increasing  the  temperature  or  doing  external  work. 
Therefore  in  the  case  of  a  perfect  gas 

dH  =  dS  +  dW 
CH*          rs2          rv, 
.'  .    (dH  =    [dS  +      dW. 

JHi  JSi  Jvi 

Integrating 

#2  -  Hi  =  S2  -  Si  +    ]  pdv 

?  rn 

Let  H  =  #2  -  #1,  S  =  S2  -  Si  and  W  =    I  pdv 

Jvi 

Then 

H  =  S  +  W.  (10) 

and  a  change  in  internal  energy  is  indicated  by  a  change  in 
temperature  alone. 

16.  Joules  Law.  When  a  perfect  gas  expands  without  doing 
external  work  and  without  taking  in  or  giving  out  any  heat,  its 
temperature  remains  unchanged  and  there  is  no  change  in  its 


ELEMENTARY  THERMODYNAMICS 


15 


internal  energy. — This  law  was  established  by  the  following  ex- 
periment performed  by  Joule. 

Two  vessels  a  and  b,  Fig.  1,  connected  by  a  tube  containing  a 
stop-cock  c  were  placed  in  a  water-bath.  One  vessel  contained 
air  compressed  to  a  pressure  of  22  atmospheres,  while  a  vacuum 
was  maintained  in  the  other.  After  the  vessels  had  remained  in 
the  bath  long  enough  so  that  the  air  and  water  were  at  the  same 
temperature  and  there  could  therefore  be  no  further  flow  of  heat 
from  one  to  the  other,  the  stop-cock  c  was  opened  and  the  air 
allowed  to  flow  from  one  vessel  to  the  other  until  the  pressure  in 
each  was  11  atmospheres.  The  temperatures  of  the  air  and 


FIG.  1. — Joule's  apparatus. 

water  were  then  read  again  and  found  to  be  unchanged.  From 
the  conditions  of  the  apparatus  no  work  external  to  the  two  vessels 
could  have  been  done.  As  the  gas  had  done  no  work  and  had 
neither  gained  nor  lost  any  heat,  its  internal  energy  must  have 
remained  unchanged.  Although  the  pressure  and  volume  of  the 
gas  had  changed  the  temperature  had  not,  thus  proving  that  a 
change  in  internal  energy  depends  upon  a  change  in  temperature 
only. 

17.  Relation  of  Specific  Heats. — If  one  pound  of  a  perfect  gas 
is  heated  at  a  constant  pressure  from  a  temperature  Ti  to  a 
temperature  T2,  and  the  volume  is  changed  from  a  volume  Vi 
to  a  volume  v2,  the  heat  absorbed  would  equal 

KP(T,  -  ro  (ii) 

and  the  work  done, 

W  =    \  pdv. 
Integrating  between  limits 


w 


r»  /•», 

=    I  pdv  =  p  I  dv  =  p(v2  - 

Jvi  Jv\ 


16  HEAT  ENGINES 

Since  from  the  equation  of  a  perfect  gas 

pv2  =  RT2,  and  pv\  = 
substituting  these  values  in  the  above  expression  for  the  work 
done,  we  have 

p(v2  -  vj  =  R(T,  -  7\).  (12) 

Since  from  equation  (10),  S  =  H  ~  W,  then  the  difference 
between  equation  (11)  and  equation  (12)  would  be  the  heat 
which  goes  to  increasing  the  temperature,  which  equals 

(Kp  -  R)  (T,  -  !Ti).  (13) 

If  the  gas  is  heated  at  a  constant  volume  from  a  tempera- 
ture Ti  to  a  temperature  T2,  then  the  heat  added  would  be 

KV(T,  -  T,),  (14) 

and  as  no  external  work  is  done  this  heat  all  goes  to  increasing 
the  temperature.  But  since  equation  (13)  also  represents  the 
heat  which  goes  to  increasing  the  temperature,  equations  (13) 
and  (14)  are  equal  to  each  other,  or 

(Kp  -  R)  (T2  -  TO  =K9(T*  -  TO, 
therefore 

Kv  =  Kp  -  R,  (15) 

or 

R  =  KP  -  Kv.  (16) 

The  difference  between  the  two  specific  heats,  R,  is  the  amount 
of  work  in  foot-pounds  done  when  one  pound  of  a  gas  is  heated 
one  degree  Fahrenheit  at  constant  pressure. 

7^ 

The  ratio  of  the  two  specific  heats,  that  is  -^,  is  denoted 

/Vtf 

by  7.  K 

Since  Kp  —  Kv  =  R,  and  J*  =  7, 

A.  v 

then 

Xp  _  R 

Kv  Kv' 

or 

R 

^"^^ 
and  hence 

K,  =  -^  (17) 

Similarly 

K,  =  (18) 


ELEMENTARY  THERMODYNAMICS 


17 


For  air 


R  =  184.77  --  131.40  =  53.37  (compare  equation  6) 


and 


K 


184.77 
131.40 


1.406. 


(19) 


18.  Expansions  in  General. — When  air,  steam,  or  any  other  gas 
is  'Used  as  the  working  substance  in  an  engine,  the  gas  is  allowed 
to  expand,  doing  work  for  a  portion  of  the  working  stroke  of 
the  engine.  The  variation  in  pressure  and  volume  during  this 
expansion  may  be  graphically  represented  by  a  mathematical 
curve  on  the  pressure-volume  plane.  The  same  is  true  in  the 
compression  of  these  gases.  On  this  plane  the  ordinates  of  any 
curve  represent  pressures  and  the  abscissae  represent  volumes. 


FIG.  2.— Paths  of  an  expanding  gas. 

Almost  all  the  expansion  or  compression  curves  ordinarily 
occurring  in  steam,  or  gas  engines,  or  the  various  forms  of 
compressors,  can  be  represented  by  the  equation 


pv 


a  constant. 


(20) 


During  expansion,  or  compression,  n  in  equation  (20)  may  have 
any  value  between  zero  and  infinity,  but  is  constant  for  any  given 
curve.  Fig.  2  shows  how  the  path  of  a  gas  will  vary  during  expan- 
sion depending  upon  the  value  of  n. 

The  value  of  n  will  determine  whether  heat  must  be  added, 
rejected,  or  remain  constant,  and  whether  the  temperature  will 
rise,  fall,  or  remain  constant  during  the  expansion,  or  compres- 
sion, of  a  gas.  These  varying  conditions  are  clearly  shown  in 
Table  V. 


18 


HEAT  ENGINES 


TABLE  V. — HEAT  AND  TEMPERATURE  CHANGES  DEPENDENT  UPON  VALUE 
OF  n  DURING  EXPANSION 


Value  of  n 

Equation  of  path 
of  gas 

Path  as 
shown  in 
Fig.  2 

Heat 

Temperature 

n  =  0  . 

p  =  constant 

ab 

Added 

1  Rises 

n  >  0  and    <  1  
n  =  1  

pvn  =  constant 
pv  =  constant 

ac 
ad 

Added 
Added 

Rises 

Constant 

n  >  1  and  <  7  

pD*  =  constant 

ae 

Added 

Falls 

n  =  y  

pifY  =  constant 

af 

Constant 

Falls 

n  >  y  and  <   °°  
n  =   oo  

pvn  —  constant 
v  =  constant 

ag 
ah 

Rejected 
Rejected 

'  Falls 
I  Falls 

For  any  path  lying  between  ad  and  a/,  heat  is  added  and  yet 
the  .temperature  falls.  In  other  words  the  specific  heat  is  nega- 
tive. 

In  case  the  gas  is  being  compressed  instead  of  expanding,  the 
changes  in  heat  and  temperature  will  be  just  the  opposite  of  those 
shown  in  the  table. 

19.  Work  of  Expansion. — The  curve  ab  in  Fig.  3  represents 
graphically  the  relation  between  pressure  and  volume  during 
expansion.     Let  the  equation  of  this  curve  be 
pvn  =  a  constant. 

In  this  figure  pressures  are  represented  by  ordinates  and  volumes 
by  abscissae.  The  gas  expands  from  a  point  a,  where  the  pres- 
sure is  pi  and  the  volume  v\,  to  the  point  b  where  the  pressure  is 
pz  and  the  volume  vz.  The  area  abed  represents  the  work  done 
during  this  expansion. 
Let  W  equal  the  work  done  during  expansion.  Then  as 


W 


r2 
pdv. 


(21) 


Since  every  point  in  the  curve  must  fulfil  the  original  conditions 
for  the  equation  of  the  curve, 

,  hence 


pvn  = 


Substituting  this  expression  in  equation  21 


,  r 

I 
•/Mi 


(22) 


(23) 


ELEMENTARY  THERMODYNAMICS 

Integrating,  W  =  piVin  - 

1  —  n 

Multiplying  out  the  parenthesis,  we  have 

w  =  PlPl  *  ^- 


l-n 


l-n 


FIG.  3. — Pressure-volume  diagram  of  an  expanding  gas. 
But  —  =  r,  the  ratio  of  expansion  for  the  gas, 

therefore 

piVi  (I  —  rl~n) 
W  =       ^j-y— 

or  substituting  p2^2n  for  piv^  in  equation  (25),  we  have 


n-l 


19 

(24) 

(25) 


W  = 


(26) 


(27) 


20  HEAT  ENGINES 

Substituting  for  pv  its  value  in  terms  of  R  and  T}  equation 
(27)  becomes 

•^:=— JiTi— •  C28) 

If  w  pounds  of  the  gas  is  expanded,  then  equation  (28)  becomes 

=  te#JTi_2\)t 

n  —  1 

20.  Heat  Added — General  Case. — In  the  case  of  any  expan- 
sion, the  heat  added  is  equal  to  the  algebraic  sum  of  heat 
equivalent  of  the  work  done  and  the  change  in  internal  energy. 
As  has  been  previously  shown,  the  change  in  internal  energy  of  a 
gas  depends  upon  the  change  in  temperature  only  and  is  equal  to 
the  heat  necessary  to  change  the  temperature  at  constant 
volume. 

Therefore  in  the  case  of  a  perfect  gas 
H  =  S  +  W 


K  ,m  m\     i  /om 

wKv(1<t  —  li)  +  -  (30) 

li    -  1 

Kv 


I 


7  -1  n-  I 


This  result  will  be  expressed  in  foot-pounds  since  pi  and  p- 
are  expressed  in  pounds  per  square  foot,  and  v\  and  vz  are  ex2 
pressed  in  cubic  feet.  To  find  the  equivalent  B.T.U.,  divide 
equation  (31)  by  778. 

Equation  (30)  may  also  be  changed  to  read  as  follows:  — 


,,,JT  (T  T  ^  _L 

~  wKv(l  2  —  L  i)  H  _ 


T,  -  T,)  +  wKv(*Y 
H  =wK,(T1-Ts)(y~    |  -1  (32) 

\  '  v    —    -L 


ELEMENTARY  THERMODYNAMICS  21 

The  answer  in  this  case  is  also  expressed  in  ft.-lbs.  since  Kv 
represents  the  specific  heat  when  expressed  in  ft.-lbs.,  and,  as  in 
equation  (31),  the  result  must  be  divided  by  778  to  find  the 
equivalent  B.T.U. 

In  equations  (30),  (31)  and  (32),  pi,  vi  and  TI  refer  to  the 
original  state  of  the  gas  and  p%,  Vz  and  Tz  to  the  final  state. 

Example.—  Five  cubic  feet  of  air  under  a  pressure  of  75  Ibs.  per  square 
inch  are  expanded  to  25  Ibs.  per  square  inch  along  a  curve  the  equation  of 
which  is  pt;1-2  =  a  constant. 

(a)  Find  the  final  volume  of  the  air.  (b)  Find  the  work  in  foot 
pounds  done  during  the  expansion,  (c)  Find  the  heat  in  B.T.U.  ,  added 
during  the  expansion. 

Solution.  —  (a)  From  equation  (20) 


Therefore,,-  g  J  ^  ^^  X  - 

=  2.26  X  51'2 
1.2  log  i'2  =  log  2.26  +  1.2  log  5 

=  .354  +  1.2  X  .699  =  .354  +  .839 
1.2  log  1-2  =  1.193 
log  r2  =  .994 

ro  =  9.86  cu.  ft. 

(b)  From  equation  (27) 


il—l 

89.7  X  144  X  5  -  39.7  X  144  X  9.86 
1.2  -  1 

64650  -56370       8280 


_  _ 

.2  .2 

=  41400  ft.-lbs. 
(c)  From  equation  (31) 


=  (8.9.7  X  144  X  5  -  39.7  X  144  X  9.86)  (1-2  _  i  ~  1.406  -  l) 


22 


HEAT  ENGINES 


=  8280 


(-  -  -  M 

\.2       .406/ 


8280  (5-2.463)  -  8280  X  2.537 


21000  ft.-lbs. 
27  B.T.U. 


21.  Expressions  for  Heat  added  at  Constant  Volume  and  at 
Constant  Pressure. — The  heat  added  at  constant  volume  may 
be  determined  from  the  volume  and  pressure,  when  the  tem- 
perature is  not  given,  in  the  following  manner: 


P2v2  T3 


PiV2T4 


FIG.  4. — Pressure-volume  diagram  when  heat  is  added  to  a  gas  at  constant 
volume  and  at  constant  pressure. 

Let  aHb  represent  the  heat  added  along  the  line  aby  Fig.  4 
Then 

aHb  =  cvw  (T2  -  TJ  in  B.T.U., 


But 


=  Kvw  (T2  -  Ti)  in  ft.-lbs. 


m  l 

wT2  =     n    and 
It 


(33) 


Substituting  these  values  in  equation  (33),  we  have 

Kvv\ 

aHb    =     -^-    (PZ    -    Pi). 

K  1 

But  from  equation  (17),  -jr  =  f 

zL  *V   """"*     J. 

Hence  substituting  in  equation  (34), 


(34) 


„ 

nb  = 


_ 


expressed  in  B.T.U. 


(35) 


ELEMENTARY  THERMODYNAMICS  23 

In  the  same  manner  we  may  derive  the  following  expression  for 
the  heat  added  at  a  constant  pressure, 

bHc  =  /^N  -»  expressed  in  B.T.U.  (36) 


Example.  —  Suppose  that  in  Fig.  4,  pi  =  151bs.  per  square  inch  abso- 
lute, pz  =  75  Ibs.  per  square  inch  absolute,  v\  =  5  cu.  ft.,  and  vz  =  25 
cu.  ft.  (a)  Find  the  heat  added  in  B.T.U.  (6)  Find  the  heat  re- 
jected in  B.T.U. 

Solution.  —  (a)  Heat  added  =  HI  =  aHb  +  bHc. 
From  equation  (35), 

77          f  i  (P2  -  Pi) 
~  (7  -  1)  X  778 

5  (75  -  15)  X  144       5  X  60  X  144      43200 
(1.406  -  1)  X  778   =      .406X778          316 

=  136.7  B.T.U. 
From  equation  (36)  . 


77         Pa  (^2  — 


(7-1)  X778 

75  X  144  (25  -  5)  X  1.406       75  X  144  X  20  X  1.406 

(1.406  -  1)  X  778  .406  X  778 

303700 
316" 

#1  =  136.7  +  961  =  1097.7  B.T.U. 
(b)  Heat  rejected  =  //2  =  cHd  +  rfffa. 

rr  ^2  (p2  ~  Pi) 

~  (7  -  1)  X"778 

=  25  (75  -  15)  X  144  _  25J*  6<^XJ44      216000 
=   (1.406"-  1)  X  778  =       .406  X  778  316 

=  683  B.T.U. 

n         P^  (^2 

(7  -  1)  X  778 

15  X  144(25-5)  X  1.406  _  15  X  144  X  20  X  1.406 
(1.406  -  1)  X  778  .406  X  778 

60740 
"319" 

#2  =  683  +  192  =  875  B.T.U. 


24  HEAT  ENGINES 

22.  Adiabatic  Expansion. — Adiabatic  expansion  is  one  in 
which  the  expanding  gas  does  not  receive  or  reject  any  heat  except 
in  the  form  of  external  work.  That  is,  there  is  no  radiation  or 
conduction  of  heat  to  or  from  the  expanding  gas,  and  the  ex- 
ternal work  is  done  at  the  expense  of  the  internal  energy  in  the 
gas.  If  compressed  adiabatically,  the  work  done  upon  the 
gas  goes  to  increasing  its  internal  energy.  Since  any  change 
in  the  internal  energy  of  a  gas  depends  upon  a  change  in  tem- 
perature, it  is  impossible  to  have  an  increase  in  the  internal 
energy  without  an  increase  in  temperature,  or  a  decrease  in 
internal  energy  without  a  decrease  in  temperature. 

Adiabatic  expansion  could  only  be  produced  in  a  cylinder 
made  of  a  perfectly  non-conducting  material  with  the  working 
fluid  itself  undergoing  no  chemical  change.  In  actual  engines, 
or  compressors,  this  is  never  the  case,  and  adiabatic  expansion 
is  only  approximated. 

Taking  the  expression 

R  (T,  -  T2) 


W  = 


n  -  1 


we  have  now  to  find  the  value  of  n  for  adiabatic  expansion. 
In  paragraph  17  it  was  shown  that  the  loss  of  energy  due  to 
a  change  of  temperature  equals 

Kv  (T,  -  7\), 
or  expressed  in  B.T.U., 

cv(T2  -  Ti). 

Equation  (10),  paragraph  15,  is 

H  =  S  +  W. 

In  adiabatic  expansion  no  heat  is  absorbed  or  rejected,  hence  H 
becomes  zero  and  W  =  —  S.  That  is,  all  the  heat  lost,  due 
to  a  change  in  temperature,  goes  to  doing  work.  (It  must 
be  understood  that  the  negative  sign  before  S  does  not  mean 
negative  work,  but  does  mean  a  decrease  in  internal  energy.) 

But  S  =  KV(TZ  -  Ti); 

therefore  the  work  done, 

W  =  KV(T,-  !F2);  (37) 

but 

R 


ELEMENTARY  THERMODYNAMICS  25 

and  hence  substituting  this  value  in  (37)  we  have 

tfCTv-jy 

7—1 

Comparing  equations  (28)  and  (38),  both  of  which  express  the 
value  for  work  done  in  an  adiabatic  expansion,  we  see  that 
n  =  7.  Therefore  the  equation  for  adiabatic  expansion  is 

pvy  =  pn) i    =  pzV2y  =  a  constant.  (39) 


Example. — Five  cubic  feet  of  air  under  a  pressure  of  75  Ibs.  per  square 
inch  are  expanded  adiabatically  until  the  pressure  is  25  Ibs. 

(a)  Find  the  final  volume  of  the  air.  (b)  Find  the  work  done  during 
the  expansion.  (See  example,  paragraph  20). 

Solution. — (a)  From  equation  (39), 

or 

v<?  =     Vi7. 
Therefore 


1.406  log  v2  =  log  2.26  +  1.406  log  5 

=  .354  +  1.406  X  .699  =  .354  +  .983 
1.406  log  v,  =  1.337 
log  V"  =  .95 

v2  =  8.915  cu.  ft. 

(b)  From  equations  (38)  and  (4), 


w  = 

7  —  1 

89.7  X  144  X  5  -  39.7  X  144  X  8.915 

1.406  -  1 
64650  -  50950  _  13700 

.406  =  .406 

=  33700  ft.-lbs. 

23.  Isothermal  Expansion.  —  A  gas  expands  or  contracts  iso- 
thermally  when  its  temperature  remains  constant  during  a  change 
of  volume.  Since  the  temperature  remains  constant  during 
isothermal  expansion  no  heat  is  absorbed  in  increasing  the 
temperature,  and  in  the  case  of  a  perfect  gas,  S  in  equation(lO) 
becomes  zero  and  H  equals  W,  or  all  the  heat  absorbed  during 
isothermal  expansion  of  a  perfect  gas  goes  to  doing  external 


26  HEAT  ENGINES 

work.     Hence  for  isothermal  expansion,  equation  (4)  becomes 
pv  =  a  constant  (40) 

(which  is  the  equation  of  a  rectangular  hyperbola)  . 

Equation  (40)  is  of  the  same  form  as  equation  (20),  and  the 
exponent  n  is  in  this  case  equal  to  1.  Substituting  1  for  the 
value  of  n  in  equation  (28),  we  derive  an  indeterminate  expres- 
sion. In  order  to  derive  an  expression  for  the  work  done  in  iso- 
thermal expansion  it  is  necessary  therefore  to  proceed  differently. 

Assume  the  curve  ab,  Fig.  3,  to  be  an  isothermal  curve,  or 
an  equilateral  hyperbola.  The  work  done  by  the  gas  in  expand- 
ing isothermally  from  volume  Vit  represented  at  the  point  a, 
to  the  volume  v2,  represented  at  the  point  6,  is  the  area  abed; 

rv2 
or  W  =    I  pdv.  (41) 

.'«1 

To  integrate  this  expression  the  pressure  must  be  expressed  in 
terms  of  volume.     From  equation  (40)  we  have 


(42) 


=  pv; 
hence 


Substituting  equation  (42)  in  equation  (41)  we  have 

W  =  piwi  j     -^. 
Integrating, 

W   =  PiVi  (loge  V2  —  log*  Vi)'} 

hence 


W  =  pivilog,.  (43) 


Since  piVi  =  RT,  then 


but       —  =  r,  the  ratio  of  expansion, 

and  piVi  =  pv; 

hence  W  =  RT  loge  r  (44) 

=  pv  \oge  r.  (45) 

If  w  pounds  of  gas  is  expanded,  then  equation  (44)  becomes 
W  =  wRT  log,  r.  (46) 


ELEMENTARY  THERMODYNAMICS  27 

During  the  isothermal  expansion  there  is  no  change  in  the 
internal  energy,  since  the  temperature  remains  constant.  Hence 
the  gas  takes  in,  during  isothermal  expansion,  an  amount  of 
heat  equivalent  to  the  work  done  during  the  expansion.  Equa- 
tions (44),  (45),  and  (46)  then  represent  not  only  the  work  done, 
but  the  equivalent  amount  of  heat  taken  in  or  rejected  during 
isothermal  expansion  or  compression. 

In  actual  practice,  when  gas  is  suddenly  compressed,  the 
compression  curve  is  approximately  an  adiabatic,  and  when 
slowly  compressed  may  be  approximately  isothermal. 

Example. — If  in  the  example  given  in  paragraph  20,  the  air  expands 
iso thermally,  find  (a)  the  final  volume  of  the  air;  (6)  the  work  done 
in  foot-pounds;  (c)  the  heat  added  in  B.T.U. 

Solution. — (a)  From  equation  (40), 

PlVi   =   P2V2 
PlVl 

or  Vz  = 

Pi 

cq  7  \/  1 44 
Therefore  v2  =  '39  7  ^  [44  X  5  =  2'26  X  5 

y2  =  11.30  cu.  ft. 
(b)  From  equation  (45), 

W  =  pivi  loge  r, 


but  r  =  *2  =  "^  =  2.26. 

Vi  O 

Therefore 

W=  89.7  X  144  X  51oge2.26 
=  89.7  X  144  X  5  X  2.3  X  .354 
=  54,200  ft.-lbs. 

(c)  Heat   added  =    --g- 

=  69.5  B.T.U. 

24.  Relation  between  p,  v,  and  T  during  Expansion  or 
Compression. — Since  the  equation  of  a  gas  during  expansion  or 
compression  is 

pvn  =  a  constant,  (see  equation  20). 
then 

piVin  =  pzVz",  (47) 

(48) 
Pi      '"  • 


28  HEAT  ENGINES 

:; 

From  the  equation  of  a  perfect  gas, 


Multiplying  equation  (47)  by  (50),  we  have 


m 

Substituting  in  equation  (51)  the  value  of  —  in  terms  of  p\  and 

v% 

p2  from  equation  (49),  we  have 


Equations  (48),  (49),  (51),  and  (52)  give  the  relations  be- 
tween pressure,  volume,  and  temperature  in  any  expansion 
or  compression. 

In  the  case  of  adiabatic  expansion  or  compression,  the  value 
of  n  in  these  equations  become  7,  (see  paragraph  22). 

Example.  —  Five  cubic  feet  of  air  under  a  pressure  of  75  Ibs.  per  square 
inch  and  at  60°  F.  are  expanded  adiabatically  until  the  pressure  is  25  Ibs. 
(See  example,  paragraph  20.)  Find  the  temperature  at  the  end  of 
expansion. 

Solution.  —  Since  the  expansion  is  adiabatic,  equation  (52)  becomes 


T~i  =  W   T 


=  (60  -f  460)   (gg'y  x  i^)1'4^  '   =  520  X  .442 

log  772  =  log  5^0  +  .29  log  .442 

=  2.716  +  .29  X  1.645  =  2.716  -  .104 
log  T2  =  2.612 
772  =  410°  abs. 

=  410  -  460  =  -  50°  F. 


ELEMENTARY  THERMODYNAMICS  29 

25.  Heat  Engine.— Any   device   used   to   convert   heat  into 
work  is  called  a  heat  engine.     The  ideally  perfect  heat  engine 
would  convert  all  the  heat  which  it  receives  into  useful  work, 
but  this  can  never  be  the  case.     This  conclusion  follows  from 
the  second  law  of  thermodynamics,  i.e.,  that  all  the  heat  which 
is  received  by  the  engine  cannot  be  converted  into  useful  work. 
In  fact  a  major  portion  of  it  is  rejected.     The  ratio  of  the  useful 
work  done  to  the  heat  received  is  called  the  heat  efficiency  of  the 
engine,  or 

Heat  equivalent  of  the  work  done 

The  heat  taken  in  by  the  engine" 

In  every  heat  engine  there  must  be  a  working  medium  for 
transferring  the  heat.  The  working  substance  may  be  solid, 
liquid,  or  gaseous.  In  all  of  the  commercial  heat  engines  now 
in  use  the  working  substance  is  a  gas.  In  the  theoretical  engine 
the  working  substance  is  supposed  to  go  through  a  cycle  of 
changes,  returning  to  its  original  condition  at  the  end  of  the 
cycle.  Each  working  cycle  involves:  first,  taking  in  the  heat 
of  the  working  substance;  second,  the  doing  of  work  by  the 
working  substance;  and  third,  the  rejection  of  heat  by  the 
working  substance.  For  example,  take  a  condensing  steam 
plant  including  the  boiler.  Water  is  fed  into  the  boiler  from 
the  hot  well  of  the  condenser.  In  the  boiler  the  water  receives 
heat  from  the  coal  and  is  transformed  into  steam.  The  steam 
carries  the  heat  to  the  engine,  part  of  which  heat  is  used  in 
doing  useful  work,  the  balance  being  lost  when  the  steam  is 
condensed  in  the  condenser.  The  condensed  steam  is  dis- 
charged into  the  hot  well  and  the  cycle  is  completed.  In  this 
cycle  of  operations  the  following  equation  must  hold  good: 

Heat  taken  in — Heat  rejected  =  Heat  equivalent  of 

work  done.  (54) 

26.  Carnot  Cycle. — The  most  efficient  means  for  converting 
heat  into  work  for  any  given  difference  in  the  temperatures  of 
the  heat  taken  in  and  heat  rejected  was  first  described  by  the 
French  engineer,  Sadi  Carnot,  in  1824. 

In  the  cycle  as  described  by  him  the  gas  first  expands  isother- 
mally  as  from  A  to  B  in  Fig.  5,  then  expands  adiabatically  from 
B  to  C,  is  then  compressed  isothermally  from  C  to  D,  and  is 
finally  compressed  adiabatically  from  D  to  A. 

To  understand  more  clearly  the  action  of  an  engine  working 


30 


HEAT  ENGINES 


in  this  cycle,  imagine  a  hot  body,  H,  Fig.  6,  as  an  infinite  source 
of  heat  at  the  temperature  T\\  a  cold  body,  C,  at  a  temperature 
T2  which  is  lower  than  TI,  and  with  an  infinite  capacity  for  ab- 
sorbing heat  without  any  change  in  temperature;  a  non-conduct- 


O  E  F  G  H 

FIG.  5. — Carnot  cycle. 

ing  cover,  N;  and  a  cylinder  covered,  except  on  the  outer  end, 
with  a  perfectly  non-conducting  material  and  containing  a  non- 
conducting, frictionless  piston.  The  outer  end  of  the  cylinder 
is  assumed  to  be  a  perfect  conductor. 


FIG.  6. — Engine  working  in  Carnot  cycle. 

The  cylinder  containing  vi  cubic  feet  of  a  perfect  gas  under 
a  pressure  pi  is  first  placed  so  that  the  conducting  end  is  in 
contact  with  the  hot  body  H  and  the  gas  allowed  to  expand  to  a 
volume  vz  and  pressure  p2.  Since  the  supply  of  heat  is  infinite, 


ELEMENTARY  THERMODYNAMICS  31 

the  temperature  will  remain  constant  and  the  expansion  will  be 
isothermal.  The  cylinder  is  next  placed  against  the  non-con- 
ducting cover,  N,  and  the  gas  allowed  to  expand  adiabatically 
to  a  volume  vs  and  pressure  ps.  At  the  same  time  the  tempera- 
ture falls  to  T2  °.  Then  the  cold  body,  (7,  is  placed  in  contact 
with  the  cylinder  head,  and  the  gas  is  compressed,  rejecting  heat 
to  C,  the  temperature  of  which  remains  constant  at  Tz,  so  that 
the  compression  is  isothermal.  Finally  the  non-conducting  cover 
is  again  placed  on  the  cylinder  head  and  the  gas  compressed 
adiabatically  to  the  original  conditions  of  pressure,  volume  and 
temperature.  The  third  step,  or  isothermal  compression  is 
carried  to  such  a  point,  D,  Fig.  5,  that  the  adiabatic  through  D 
will  pass  through  A. 

The   heat    absorbed    along    the    isothermal   AB,   Fig.   5,  is 

equal  to  —j-  loge     ,  and  the  heat  rejected  along  CD  is  equal  to 

—j-  loge  --"  As  BC  and  DA  are  adiabatics  there  will  be  no  heat 
J  v± 

received  or  rejected  along  these  lines,  and  all  the  heat  will  be 
received  along  AB  and  all  the  heat  rejected  along  CD. 

Since  the  temperature  along  AB  is  TI,  and  along  CD  is  r2, 
then 

Tl  =-  ( 
T2 
Hence 


Let  HI  equal  the  heat  added,  and  H2  the  heat  rejected,  and  let 

W 

j  equal  the  heat  equivalent  of  the  work  done.     Then 

w     77       rr 

j    —  JnLi  —  ri2, 

and  the  efficiency  is 

W        #1  -  #2 


-* 
ST 

Substituting  in  this  expression,  the  expression  for  the  heat 
absorbed  and  the  heat  rejected  as  given  above,  we  have  the 
efficiency 

v3 


32  HEAT  ENGINES 

Substituting  for  —  its  value  in  terms  of  v%  and   v\,  and  simpli- 
fying the  expression, 


E  _          -  (56) 

From  the  equation  of  a  perfect  gas, 

piVi  =  RT1}  and  psvs  =  RT2. 
Substituting  in  equation  (56), 

E  =  T^f-  (57) 

This  expression  for  efficiency  is  general  for  all  engines  using 
perfect  gases,  as  in  deriving  the  expression  we  have  not  assumed 
any  special  conditions  dependent  upon  the  nature  of  the  gas. 

Equation  (57)  can  only  be  unity  when  T2  =  0,  that  is,  when 
the  temperature  of  the  condenser,  or  cold  body,  is  absolute 
zero.  The  nearer  unity  equation  (57)  becomes,  the  higher  the 
efficiency  of  the  engine.  In  order  to  obtain  this  result,  T\ 
-  T2  must  be  made  as  large  as  possible.  This  can  only  be 
attained  by  making  T\  larger,  or  Tz  smaller.  In  actual  practice 
there  are  limits  to  the  values  of  TI  and  T%  which  may  be  available 
in  the  different  forms  of  engines. 

It  may  also  be  shown  by  the  following  demonstration  that 
in  any  working  medium  in  which  equal  increments  of  tempera- 
ture represent  equal  increments  of  heat,  the  expression  for 
efficiency  applies. 

Assume  a  scale  of  temperature  so  that  each  degree  on  the 
temperature  scale  represents  one  heat  unit,  then  heat  and  tem- 
perature would  be  represented  by  the  same  quantity  numeri- 
cally. From  equation  (55) 

Hi—  H<i 

//i    ' 

but  on  the  assumed  temperature  scale 

#!  -  Tlt  and  #2  -  T>, 

fji     _    rrt 

hence  E  =        ™         (compare  equation  57) 

i  i 

All  experience  in  testing  engines  using  either  perfect  or 
imperfect  gases  as  their  working  medium  goes  to  show  that 
this  law  applies  to  all  forms  of  engines  no  matter  what  the 
working  medium  may  be. 


ELEMENTARY  THERMODYNAMICS  33 

27.  Reversibility  of  Carnot  Cycle. — The  Carnot  cycle  is  a 
reversible  one  as  the  gas  may  be  considered  to  first  expand 
adiabatically  along  AD  and  then  isothermally  along  DC,  then 
to  be  compressed  adiabatically  along  CB,  and  finally  com- 
pressed isothermally  along  BA.  It  is  thus  possible  to  work 
around  the  cycle  in  the  reverse  direction. 

Having  proved  that  the  Carnot  cycle  is  reversible  and  that 

rp       _    np 

its  efficiency  is  equal  to  — ^r — ,  it  is  now  necessary  to  show 

that  no  cycle  can  be  more  efficient  than  a  reversible  one,  and  that 
no  reversible  cycle  can  have  a  greater  efficiency  than  that  of 
the  Carnot  cycle. 

Assume  a  non-reversible  engine  A  and  a  Carnot  engine  B, 
both  working  between  the  same  limits  in  temperature.  Engine 
A  takes  QA  heat  units  from  the  hot  body  and  rejects  Q'A  heat 
units  to  the  cold  body,  while  engine  B  takes  QB  heat  units  from 
the  hot  body  and  rejects  Q'B  heat  units  to  the  cold  body. 

If  engine  A  is  more  efficient  than  engine  B,  it  must  take  less 
heat  from  the  hot  body  and  reject  less  to  the  cold  body,  or  in 
other  words 

QA   <  QB 
and  Q'A  <  Q'B. 

Now  assume  that  B  is  to  run  in  the  reverse  direction  and 
that  A  is  to  drive  5,  which  acts  as  a  heat  pump.  Since  B  is  a 
reversible  engine,  it  will  reject  to  the  hot  body,  when  running 
in  a  reverse  direction,  the  same  amount  of  heat  that  it  takes 
from  that  body  when  running  direct.  Therefore  the  combined 
unit  of  A  —  B  will,  in  each  cycle,  take  from  the  hot  body  the 
quantity  of  heat  QA  and  reject  to  the  hot  body  the  quantity  of 
heat  QB- 
But  QB  >  QA 

which  means  that  this  "  self-acting  machine  unaided  by  any 
external  agency"  is  transferring  heat  from  a  body  of  lower  to 
one  of  higher  temperature.  This  is  contrary  to  the  Second 
Law  of  Thermodynamics.  It  is,  therefore,  impossible  for  engine 
A  to  be  more  efficient  than  engine  B.  As  these  represent  any 
engines  of  these  particular  types,  no  non-reversible  engine  can 
be  more  efficient  than  a  reversible  one  working  in  the  Carnot 
cycle. 

Now  assume  engine  A  to  be  a  reversible  engine  also.     It  can 

3 


34  HEAT  ENGINES 

be  similarly  proven  that  it  cannot  be  more  efficient  than  the 
Carnot  engine. 

The  conclusion  is  therefore  reached  that  no  cycle  can  be 
more  efficient  than  the  Carnot  cycle. 

It  can  also  be  proven  that  this  cycle  is  the  most  efficient 
cycle  that  any  engine  can  follow  when  working  between  any 
given  temperature  limits.  This  necessitates,  however,  a  more 
thorough  exposition  of  the  principles  of  thermodynamics  than 
it  is  deemed  wise  to  include  in  this  text,  and  will  therefore  be 
omitted. 

PERFECT  GAS  PROBLEMS 

1.  One  pound  of  air  under  a  pressure  of  100  Ibs.  per  square  inch  absolute 
occupies  .3  of  a  cubic  foot  in  volume.     What  is  its  temperature  in  degrees  F.  ? 

2.  Ten  pounds  of  air  under  a  pressure  of  10,000  Ibs.  per  square  inch 
absolute  have  a  temperature  of  100°  F.     Find  the  volume  occupied. 

3.  Five  pounds  of  air  at  a  temperature  of  60°  F.  occupy  a  volume  of  50 
cu.  ft.     Find  the  gage  pressure  per  square  inch. 

4.  A  tank  containing  air  has  a  volume  of  300  cu.  ft.     The  pressure  in  the 
tank  is  100  Ibs.  per  square  inch  absolute  and  the  temperature  is  70°  F.     Find 
the  weight  of  air  in  the  tank. 

6.  What  is  the  weight  of  the  quantity  of  air  which  occupies  a  volume  of 
10  cu.  ft.  at  a  temperature  of  100°  F.  under  a  pressure  of  50  Ibs.  per  square 
inch  absolute? 

6.  What  is  the  temperature  of  a  pound  of  air  when  its  volume  is  5  cu.  ft. 
and  the  pressure  is  35  Ibs.  per  square  foot  absolute? 

7.  What  is  the  weight  of  a  cubic  foot  of  air  when  the  pressure  is  50  Ibs. 
per  square  inch  absolute  and  the  temperature  160°  F.? 

8.  A  quantity  of  air  at  a  temperature  of  60°  F.  under  a  pressure  of  14.7 
Ibs.  per  square  inch  absolute  has  a  volume  of  5  cu.  ft.     What  is  the  volume 
of  the  same  air  when  its  temperature  is  changed  to  120°  F.  at  constant 
pressure? 

'"9.  The  volume  of  a  quantity  of  air  at  a  temperature  of  60°  F.  under  a 
pressure  of  14.7  Ibs.  per  square  inch  absolute  is  10  cu.  ft.  What  is  the  volume 
of  the  same  air  when  the  pressure  is  changed  at  constant  temperature  to 
60  Ibs.  per  square  inch  absolute? 

10.  A  tank  contains  200  cu.  ft.  of  air  at  a  temperature  of  60°  F.  and  under 
a  pressure  of  200  Ibs.  absolute,     (a)  What  is  the  weight  of  the  air?     (b) 
How  many  cubic  feet  will  the  air  occupy  at  atmospheric  pressure? 

11.  A  tank  containing  1000  cu.  ft.  is  half  full  of  air  and  half  full  of  water. 
The  pressure  in  the  tank  is  60  Ibs.  absolute  and  the  temperature  is  60°  F.     If 
half  the  water  is  withdrawn  from  the  tank,  what  will  be  the  resulting  pressure, 
assuming  the  temperature  to  remain  constant? 

12.  The  volume  of  a  quantity  of  air  at  70°  F.  under  a  pressure  of  16  Ibs. 
per  square  inch  absolute  is  20  cu.  ft.     What  is  the  temperature  of  this  air 
when  the  volume  is  4  cu.  ft.  and  the  pressure  is  70  Ibs.  per  square  inch 
absolute? 


ELEMENTARY  THERMODYNAMICS  35 

13.  A  compressed  air  pipe  transmission  is  1  mile  long.     The  pressure  at 
entrance  is  1000  Ibs.  per  square  inch  absolute;  at  exit,  500  Ibs.     The  velocity 
at  entrance  to  pipe,  which  is  12  in.  in  diameter,  is  100  ft.  per  second,     (a) 
What  must  be  the  diameter  of  the  pipe  at  the  exit  end  to  have  the  same 
velocity  as  at  entrance,  the  temperature  of  the  air  in  the  pipe  remaining  con- 
stant?    (&)  What,  if  the  velocity  at  exit  is  to  be  90  ft.  per'  second? 

14.  A  street  car  has  an  air  storage  tank  for  its  air  brakes  with  a  volume 
of  400  cu.  ft.     The  pressure  in  the  tank  at  starting  is  200  Ibs.  absolute  and  the 
temperature  is  60°  F.     The  air-brake  cylinders  take  air  at  40  Ibs.  absolute 
and  have  a  volume  of  2  cu.  ft.     How  many  times  can  the  brakes  be  operated 
on  one  tank  of  air,  assuming  the  temperature  of  the  air  to  remain  constant? 

15.  To  operate  the  air  brakes  on  a  car  requires  1  cu.  ft.  of  air  at  40  Ibs. 
gage  pressure.     The  car  has  a  storage  tank  containing  100  cu.  ft.  of  air  at 
250  Ibs.  gage  pressure.     How  many  times  will  the  tank  operate  the  brakes? 

The  compressed  air  tank  on  a  street  car  has  a  volume  of  250  cu.  ft. 
The  pressure  in  the  tank  is  250  Ibs.  gage  and  the  temperature  is  60°  F.  There 
are  two  air  cylinders  each  8"  X  10".  The  brakes  take  air  aT*40  Ibs.  gage 
pressure  and  60°  temperature.  How  many  times  will  the  tank  operate 
the  brakes? 

17.  How  many  B.T.U.  will  be  required  to  double  the  volume  of  1  Ib.  of  air 
at  constant  pressure  from  the  temperature  of  melting  ice? 

18.  A  tank  filled  with  200  cu.  ft.  of  air  at  atmospheric  pressure,  and  at 
60°  F.  is  heated  to  150°.     What  will  be  the  resulting  air  pressure  in  the  tank 
and  how  many  B.T.U.  will  be  required  to  heat  the  air? 

AX^^  A  tank  contains  200  cu.  ft.  of  air  at  60°  F.  under  a  pressure  of  40  Ibs. 
\ibk>lute.     If  the  air  has  1000  B.T.U.  added  to  it,  what  will  be  the  resulting 

temperature  and  pressure  in  the  tank? 

20.  A  tank  contains  100  cu.  ft.  of  air  at  60°  F.  under  a  pressure  of  50  Ibs. 

absolute.     If  the  air  in  the  tank  receives  100  B.T.U.  of  heat,  what  will  be 

the  resulting  temperature  and  pressure? 
\J    21.  Ten  pounds  of  air  enclosed  in  a  tank  at  60°  F.  under  a  pressure  of  100 

Ibs.  absolute  are  heated  to  100°  F.     (a)  What  is  the  volume  of  the  air? 

(&)  What  will  be  the  final  pressure?     (c)  How  many  B.T.U.  will  be  required 

to  heat  it? 

22.  A  tank  contains  200  cu.  ft.  of  air  at  60°  F.  under  a  pressure  of  50  Ibs. 
absolute,     (a)  Ho  w  many  pounds  of  air  in  the  tank  ?     (6)  How  many  B.T.U. 
will  be  required  to  raise  the  temperature  of  the  air  in  the  tank  to  100°  F.? 
(c)  What  will  be  the  pressure  in  the  tank  when  the  air  has  been  heated  to 
100°  F.? 

23.  A  certain  auditorium  will  seat  3000  people.     If  each  person  is  supplied 
with  2000  cu.  ft.  of  air  per  hour  for  ventilation,  the  outside  temperature 
being  0°  F.  and  that  in  the  hall  being  70°,  how  many  pounds  of  air  will  be 
admitted  per  hour,  and  how  many  B.T.U.  will  be  required  to  heat  it? 
Weight  of  1  cu.  ft.  of  air  at  0°  F.  is  .0863  Ibs.;  at  70°  is  .075  Ibs. 

24.  A  piece  of  iron  weighing  5  Ibs.  is  heated  to  212°  F.  and  then  dropped 
into  a  vessel  containing  16.5  Ibs.  of  water  at  60°  F.     If  the  temperature  of 
the  water  is  increased  five  degrees  by  the  heat  from  the  iron,  what  is  the 
specific  heat  of  the  iron? 

\/  25.  How  many  foot-pounds  of  heat  must  be  absorbed  by  2  Ibs.  of  air  in 


36  HEAT  ENGINES 

expanding  to  double  its  initial  volume  at  constant  temperature  of  100°  F.? 

26.  How  many  B.T.U.  of  work  must  be  expended  in  compressing  3  Ibs. 
of  air  to  one-fourth  its  initial  volume  at  a  constant  temperature  of  15°  C.? 

27.  If  1  cu.  ft.  of  air  expands  from  a  gage  pressure  of  4  atmospheres  and 
a  temperature  of  60°  F.  to  an  absolute  pressure  of  1  atmosphere  without  the 
transmission  of  heat,  find  the  final  temperature. 

28.  An  air  compressor,  the  cross-section  of  which  is  2  sq.  ft.,  and  stroke 
3  ft.,  takes  in  air  at  14  Ibs.  absolute  pressure  and  60°  F.  and  compresses  it  to 
60  Ibs.  gage  pressure  without  the  transmission  of  heat.     Find  the  final 
temperature. 

29.  In  problem  28,  if  the  air  at  60  Ibs.  gage  pressure  and  70°  F.  expands 
adiabatically  to  a  final  pressure  of  20  Ibs.  gage,  find  the  final  temperature. 

30.  Two  cubic  feet  of  air  at  60°  F.  and  an  initial  pressure  of  1  atmosphere 
absolute  are  compressed  in  a  cylinder  to  5  atmospheres  gage  pressure. 
If  there  be  no  transference  of  .heat,  find  the  final  temperature  and  volume. 

31.  A  cylindrical  vessel,  the  area  of  the  base  of  which  is  1  sq.  ft.,  contains 
2  cu.  ft.  of  air  at  60°  F.  when  compressed  by  a  frictionless  piston  weighing 
2000  Ibs.  resting  upon  it.     Find  the  temperature  and  volume  of  the  air  if  the 
vessel  be  inverted,  there  being  no  transmission  of  air  or  heat. 

32.  Given  the  quantity  of  air  whose  volume  is  3  cu.  ft.  at  60°  F.  under  a 
pressure  of  45  Ibs.  absolute,     (a)  Find  the  volume  and  temperature  of  this 
air  after  it  has  expanded  adiabatically  until  its  pressure  is  15  Ibs.  absolute. 
(b)  What  is  the  work  done  during  the  expansion?     (c)  What  is  the  heat  in 
B.T.U.  converted  into  work? 

33.  Given  a  quantity  of  air  whose  volume  is  2  cu.  ft.  at  60°  F.  under  a 
pressure  of  80  Ibs.  absolute,     (a)  What  is  the  weight  of  the  air?     (6)  What 
will  be  the  final  temperature  and  pressure  if  the  air  be  expanded  adiabatically 
until  its  volume  is  8  cu.  ft.?     (c)  How  much  work  will  be  done  during  this 
expansion?     (d)  How  much  work  will  be  done  if  the  air  be  expanded  isother- 
mally until  its  volume  is  8  cu.  ft.? 

34.  Given  a  quantity  of  air  whose  volume  is  2.2  cu.  ft.  at  80°  F.  under  a 
pressure  of  100  Ibs.  absolute.     It  is  made  to  pass  through  the  following 
Carnot  cycle:     it  is  expanded  isothermally  until  its  volume  is  4  cu.  ft.;  then 
expanded  adiabatically  until  its  temperature  is  30°  F.;  then  compressed 
isothermally;  and  finally  it  is  compressed  adiabatically  until  its  volume,  pres- 
sure, and  absolute  temperature  are  the  same  as  at  the  beginning  of  the  cycle. 
(a)  Find  the  total  heat  added  in  B.T.U.     (6)  Find  the  total  heat  rejected 
in  B.T.U.     (c)  Find  the  work  done  in  foot-pounds  during  the  cycle,     (d) 
Find  the  efficiency  of  the  cycle. 

35.  Given  a  quantity  of  air  whose  volume  is  10  cu.  ft.  at  60°  F.  under  a 
pressure  of  20  Ibs.  absolute.     Heat  is  added  at  constant  volume  until  its  pres- 
sure is  200  Ibs.  absolute;  then  added  at  constant  pressure  until  its  volume  is 
40  cu.  ft.;  then  rejected  at  constant  volume  until  its  pressure  is  20  Ibs.  abso- 
lute; and  then  rejected  at  constant  pressure  until  its  volume  is  the  same  as 
at  the  beginning  of  the  cycle,     (a)  Find  temperature  at  end  of  first  step. 
(6)  Find  temperature  at  end  of  second  step,     (c)  Find  temperature  at  end 
of  third  step,     (d)  Find  total  heat  added  in  B.T.U.     (e)  Find  total  heat 
rejected  in   B.T.U.     (/)  Find  work  done  in  foot-pounds,     (g)  Find  the 
efficiency  of  the  cycle. 


ELEMENTARY  THERMODYNAMICS  37 

36.  Given  a  quantity  of  air  whose  volume  is  20  cu.  ft.  at  60°  F.  under  a 
pressure  of  20  Ibs.  absolute.     Heat  is  added  at  constant  volume  until  its 
pressure  is  200  Ibs.  absolute;  then  the  air  is  expanded  adiabatically  until  its 
pressure  is  20  Ibs.  absolute;  and  then  compressed  at  constant  pressure  until 
its  volume  is  the  same  as  at  the  beginning  of  the  cycle,     (a)  Find  tempera- 
ture at  end  of  first  step,     (b)  Find  temperature  at  end  of  second  step,     (c) 
Find  total  heat  added  in  B.T.U.     (d)  Find  total  heat  rejected  in  B.T.U. 

37.  Given  a  quantity  of  air  whose  volume  is  1  cu.  ft.  under  a  pressure  of 
100  Ibs.  absolute.     It  is  expanded  under  a  constant  pressure  to  3  cu.  ft.     (a) 
What  external  work  has  been  done  during  the  expansion?     (6)  What  heat 
has  been  added? 

•  38.  Two  pounds  of  air  occupying  a  volume  of  6  cu.  ft.  under  a  pressure 
of  60  Ibs.  absolute  are  expanded  isothermally  until  the  pressure  is  20  Ibs. 
absolute,  (a)  What  external  work  has  been  done  during  the  expansion? 
(6)  What  heat  has  been  added? 

39.  1.3  cu.  ft.  of  air  under  a  pressure  of  15  Ibs.  absolute  are  heated  at 
constant  volume  to  80  Ibs.  absolute;  then  expanded  adiabatically  to  a  volume 
of  4.26  cu.  ft.  and  a  pressure  of  15  Ibs.  absolute;  then  compressed  at  a  con- 
stant pressure  to  the  original  volume,     (a)  What  is  the  total  heat  added  in 
B.T.U.?     (6)  What  is  the  Work  done  in  foot-pounds?     (c)  What  is  the 
efficiency  of  the  cycle? 

40.  Two  cubic  feet  of  air  under  a  pressure  of  15  Ibs.  per  square  inch  abso- 
lute are  heated  at  constant  volume  to  a  pressure  of  100  Ibs.  per  square  inch 
absolute;  then  heated  at  constant  pressure  to  a  volume  of  4  cu.  ft.;  then 
expanded  to  the  original  pressure;  and    finally  compressed  at  constant 
pressure  to  the  original  volume.     The  expansion  is  pv1-2  =  a  constant,     (a) 
Find  the  heat  added  in  B.T.U.     (6)  Find  the  heat  rejected  in  B.T.U.     (c) 
Find  the  work  done  in  foot-pounds,     (d)  Find  the  efficiency  of  the  cycle. 
'  (41,  One  pound  of  air  is  made  to  pass  through  the  following  cycle:  it  is 
expanded  at  constant  pressure;  then  expanded  isothermally;  then  com- 
pressed at  constant  pressure;  and  then  compressed  isothermally  until  the 
cycle  is  complete.     Derive  the  expressions  in  terms  of  pressure  and  volume 
for,  (a)  the  heat  added  in  B.T.U.;  (6)  the  heat  rejected  in  B.T.U.;  (c)  the 
work  done  in  foot-pounds;  (d)  the  efficiency  of  the  cycle. 

<*»       /  /   /  yr     WIA     /  /i/L/r  ^  ^      ~T>     /is   -ri/ 


f  '*  £ 

\   >^<&yvi/t 

tjv  & 

// N  £  3-f fa 


./ 


CHAPTER  III 
PROPERTIES  OF  STEAM 

28.  Formation  of  Steam. — In  order  to  understand  the  opera- 
tion of  a  steam  engine  it  is  necessary  to  study  the  nature  and 
properties  of  steam.  Steam  as  produced  in  the  ordinary  boiler 
is  a  vapor,  and  often  contains  a  certain  amount  of  water  in 
suspension,  as  does  the  atmosphere  in  foggy  weather.  Let  us 
suppose  that  we  have  a  boiler  partly  filled  with  cold  water, 
and  that  heat  is  applied  to  the  external  shell  of  the  boiler.  As 
the  water  in  the  boiler  is  heated  its  temperature  slowly  rises. 
This  increase  of  temperature  continues  from  the  initial  tem- 
perature of  the  water  until  the  temperature  of  the  boiling  point 
is  reached,  this  latter  temperature  depending  upon  the  pressure 
in  the  boiler.  When  the  boiling  point  is  reached  small  par- 
ticles of  water  are  changed  into  steam.  They  rise*  through 
the  mass  of  water  and  escape  to  the  surface.  The  water  is  then 
said  "to  boil."  The  temperature  at  which  the  water  boils  depends 
entirely  on  the  pressure  in  the  boiler.  The  steam  produced  from 
the  boiling  water  is  at  the  same  temperature  as  the  water,  and 
under  this  condition  the  steam  is  said  to  be  saturated.  If  we 
keep  on  applying  heat  to  the  water  in  the  boiler,  the  pressure 
remaining  the  same,  the  temperature  of  the  steam  and  the 
water  will  remain  constant  until  all  the  water  is  evaporated. 
If  more  heat  is  added  after  all  the  water  is  converted  into  steam, 
the  pressure  still  being  kept  unchanged,  the  temperature  will  rise. 
Steam  under  this  condition  is  said  to  be  superheated. 

In  the  formation  of  steam  we  divide  the  heat  used  into  three 
different  parts: 

(1)  The  heat  which  goes  to  raising  the  temperature  of  the 
water  from  its  original  temperature  to  the  temperature  of  the 
boiling  point,  called  "Heat  of  the  Liquid." 

(2)  The  heat  which  goes  to  changing  the  water  at  the  tem- 
perature of  the  boiling  point  into  steam  at  the  temperature 
of  the  boiling  point,  called  "Latent  Heat." 

(3)  The  heat  which  goes  to  changing  the  saturated  steam  at 
the  temperature  of  the  boiling  point  into  steam  at  a  higher 

38 


PROPERTIES  OF  STEAM  39 

temperature  but  at  the  same  pressure,  called  "Heat  of  Super- 
heat." 

29.  Dry  Saturated  Steam. — Saturated  steam  always  exists  at 
the  temperature  of  the  boiling  point  corresponding  to  the  pres- 
sure.    If  this  saturated  steam  contains  no  moisture  in  the  form 
of  water,  then  it  is  said  to  be  dry  saturated  steam,  or,  in  other 
words,  dry  saturated  steam  is  steam  at  the  'temperature  of  the 
boiling  point  and  containing  no  water  in  suspension.     Water 
so   contained   is   often   called   entrained   moisture.     If   heat  is 
added  to  dry  saturated  steam,  not  in   the  presence  of  water 
it  will  become  superheated.     If  heat  is  taken  away  from  dry 
saturated   steam   it   will   become   wet   steam.     Dry   saturated 
steam  is  not  a  perfect  gas,  and  the  relation  of  pressure,  volume, 
and  temperature  for  such  steam  does  not  follow  any  simple 
law,  but  has  been  determined  by  experiment. 

The  properties  of  dry  saturated  steam  were  originally  deter- 
mined by  Regnault  between  sixty  and  seventy  years  ago,  and 
so  carefully  was  his  work  done  that  no  errors  in  his  results  were 
apparent  until  within  very  recent  years,  when  the  great  diffi- 
culty in  obtaining  steam  which  is  exactly  dry  and  saturated 
became  appreciated,  and  new  experiments  by  various  scientists 
proved  that  Regnault' s  results  were  slightly  high  at  some  pres- 
sures and  slightly  low  at  others.  The  steam  tables  given  in 
this  book  are  based  upon  these  recent  experiments,  and  are 
probably  correct  to  a  fraction  of  1  per  cent. 

30.  Wet   Steam. — Wet  steam  is  saturated  steam  which  con- 
tains entrained  moisture.     When   saturated   steam  is   used   in 
a  steam  engine,  it  almost  always  contains  moisture  in  the  form 
of  water,  so  that  the  substance  used  by  the  engine  as  a  working 
fluid  is  a  mixture  of  steam  and  water.     The  steam  and  water 
in  this  case  are  at  the  same  temperature. 

31.  Superheated   Steam. — Superheated   steam  is   steam   at  a 
temperature    higher    than    the    temperature    corresponding    to    the 
pressure  of  the  boiling  point  at  which  it  was  formed.     It  is  some- 
times called  steam  gas.     If  water  were  to  be  mixed  with  super- 
heated steam,  this  water  would  be  evaporated  as  long  as  the 
steam  remains  superheated.     Superheated  steam  at  the  same 
pressure  as  the  boiling  point  at  which  it  was  produced  can  have 
any  temperature  higher  than  that  of  the  boiling  point.     When 
raised  to  any  considerable  temperature  above  the  temperature 
of  the  boiling  point,  it  follows  very  closely  the  laws  of  a  perfect 


40 


HEAT  ENGINES 


gas,  and  may  be  treated  as  a  perfect  gas.     The  equation  for 
superheated  steam,  considered  as  a  perfect  gas,  is 
pv  =  85.5  T,  approximately. 

The  specific  heat  of  superheated  steam  is  a  variable  and 
depends  upon  the  pressure  of  the  steam  and  the  temperature  to 
which  the  steam  is  superheated.  For  approximate  calculations, 
the  following  values  for  the  specific  heat  of  superheated  steam 
may  be  taken. 

TABLE  VI.  SPECIFIC  HEATS  OF  SUPERHEATED  STEAM 


Abs.  press,  in 

14.7    25.0    50.0    75.0100.0125.0150.0175.0200.0225.0250.0275.0300.0 

Ibs.  per  sq.  in. 

Temp,  of  boil- 
ing point,  F.° 

212.0 

240.1  281.0  307.6  327.8  344.4  358.5 

370.8381.9  391.9401.1 

409.5417.5 

Actual  Temp. 

I 

of  Steam. 

250° 

47 

48 

275 

.47 

.48 

I 

300 

47       a.« 

50 

325 

47 

4.8 

50        53 

350 

47       48 

49        *>9 

55 

58 

375 

.47      .47 

.49 

.51 

.53 

.56 

.60 

.66 

400 

.47      .47 

.49 

.50 

.52 

.54 

.57 

.60 

.65 

.72 

425 

.47      .47 

.48 

.50 

.51 

.52 

.54 

.56      .59 

.63 

.67 

.74 

.82 

450 

.47 

.47 

.48      .49 

.50 

.51 

.52 

.53      .55 

.57 

.60 

.64 

.67 

475 

.47      .47 

.48      .49 

.50 

.50 

.51 

.52 

.53 

.54 

.55 

.56 

.58 

500 

.47      .47 

.48 

.49 

.49 

.50 

.50 

.51 

.52 

.52 

.53 

.53 

.54 

525 

.47      .47 

.48      .48 

.49 

.49 

.50 

.50 

.51 

.51 

.51 

.52 

.52 

550 

.47      .47 

.48      .48 

.49 

.49 

.50 

.50 

.50 

.50 

.51 

.51 

.52 

600 

.47      .47 

.48 

.48 

.48      .49 

.49 

.49 

.49 

.50 

.50 

.50 

.50 

650 

.47 

.47 

.48 

.48 

.48      .48 

.49 

.49. 

.49 

.49 

.50 

.50 

.50 

700 

.47 

.47 

.48 

.48 

.48      .48 

.49 

.49 

.49 

.49 

.49 

.50 

.50 

800 

.48       .48 

.48 

.48 

.48      .48   !   .48 

.49      .49 

.49 

.49 

.49 

.49 

When  more  accurate  results  are  desired  the  value  of  specific 
heat  should  be  taken  from  results  given  in  Peabody's,  or  Marks 
and  Davis's  Steam  Tables. 

The  value  of  y  for  superheated  steam  is  approximately  1.3. 

32.  Heat  of  the  Liquid. — The  heat  necessary  to  raise  one  pound 
of  water  from  32°  to  the  temperature  of  the  boiling  point  is  called 
the  heat  of  the  liquid.  This  may  be  expressed  numerically  as 
follows:  let  c  be  the  specific  heat  of  the  water,  t  the  temperature 
of  the  boiling  point,  and  h  the  heat  of  the  liquid; 
then 

h  =  c  (t  -  32).  (1) 

For  approximate  results  c  may  be  taken  as  1,  but  where  great 
accuracy  is  required  the  heat  of  the  liquid  should  be  taken  from 


PROPERTIES  OF  STEAM  41 

the  steam  tables  as  shown  in  Column  3.  During  this  operation 
the  change  in  the  volume  of  the  water  is  extremely  small,  and 
the  amount  of  external  work  done  may  be  neglected  and  all  the 
heat  of  the  liquid  may  be  considered  as  going  to  increasing  the 
heat  energy  of  the  water. 

33.  Latent  Heat  of  Steam. — When  the  water  has  reached  the 
boiling  point,  more  heat  must  be  added  to  convert  this  water 
into  steam.     The  heat  necessary  to  convert  one  pound  of  water  at 
the  temperature  of  the  boiling  point  into  steam  at  the  same  tempera- 
lure  is  called  the  latent  heat.     We  will  denote  the  latent  heat  by 
L.     Experiments  show  that  the  latent  heat  of  steam  diminishes 
as  the  pressure  increases. 

When  water  is  changed  into  steam,  the  volume  is  increased 
rapidly  so  that  a  considerable  portion  of  the  latent  heat  goes 
to  external  work.  Let  P  equal  the  pressure  at  which  the  steam 
is  formed;  V  equal  the  volume  of  the  steam,  and  v  equal  the 
volume  of  the  water:  then  the  external  work  done  equals 

P  (V  -  v).  (2) 

The  volume  of  one  pound  of  water  under  those  conditions  may 
be  taken  as  approximately  .017  cu.  ft.  At  212°  the  external 
work  done  in  producing  one  pound  of  steam  is  equivalent  to 
73  heat  units  or  about  one-thirteenth  of  the  latent  heat. 

Experiments  show  that  the  latent  heat  of  steam  diminishes 
about  .695  heat  units  for  each  degree  the  temperature  of  the 
boiling  point  is  increased.  If  t  be  the  temperature  of  the  boiling 
point,  then,  approximately, 

L  =  1072.6  -  .695  (i  -  32).  (3) 

In  condensing  steam  the  same  amount  of  heat  is  given  up 
as  was  required  to  produce  it. 

34.  Total  Heat  of  Steam. — The  total  heat  of  steam  is  the  heat 
necessary  to  change  one  pound  of  water  at  32°  to  one  pound  of 
steam  at  the  temperature  of  the  boiling  point.     The  total  heat  of 
dry  saturated  steam  will  be  designated  by  H. 

H  =  h  +  L.  (4) 

The  experimental  results  as  given  in  the  table  for  the  value  of 
the  total  heat  may  be  approximated  very  closely  by  the  formula 
H  =  1072.6  +  .305  (t  -  32).  (5) 

It  is  more  accurate,  however,  to  take  the  values  of  the  total 
heat  from  the  tables  than  it  is  to  compute  them  from  the  formula 
given. 


42 


HEAT  ENGINES 


If  we  let  q  represent  the  percentage  of  dry  steam  in  a  mixture 
of  steam  and  water,  then  the  latent  heat  in  one  pound  of  wet 
steam  equals 

qL  (6) 

and  the  total  heat  of  one  pound  of  wet  steam  equals 

h  +  qL.  (7) 

35.  Steam  Tables. — The  following  table  shows  the  properties 
of  dry  saturated  steam.  More  complete  tables  will  be  found 
in  Peabody's  Steam  Tables,  Marks  and  Davis's  Steam  Tables, 
or  in  the  Engineering  Hand  Books.  Column  1  gives  the  ab- 
solute pressure  of  the  steam  in  pounds  per  square  inch.  Column 
2  gives  the  corresponding  temperature  of  the  steam  in  Fahren- 
heit degrees.  Column  3  gives  the  heat  of  the  liquid,  or  the  heat 
necessary  to  raise  one  pound  of  water  from  32  degrees  to  the  boil- 
ing point  corresponding  to  the  pressure.  Column  4  gives  the 
latent  heat,  or  the  heat  necessary  to  change  a  pound  of  water 
at  the  temperature  of  the  boiling  point  into  steam  at  the  same 
temperature.  Column  5  gives  the  total  heat  of  the  steam,  and 
is  the  sum  of  the  quantities  in  Column  3  and  Column  4.  Column 
6  is  the  volume  of  one  pound  of  steam  at  the  different  tempera- 
tures. Column  7  is  the  weight  of  one  cubic  foot  of  steam  at  the 
different  temperatures. 

TABLE  VII. — PROPERTIES  OF  SATURATED  STEAM 

ENGLISH   UNITS 


ii. 

* 

*, 

M 

ii 

«SS^ 

1 

5  « 

in  ^  d 

grc.s 

2  » 

•si 

w  §§ 

Kg 

jc  S  _^  § 

•5  «fe 

VO™ 

II 

?5 

11^ 

l£ 

jyhi 

ils 

fift 

^ 

o  Q 

w 

In 

H° 

0 

OH 

<^ 

P 

* 

fc 

L 

/T 

0 

i 
S 

P 

.0886 

32 

0 

1072.6 

1072.6 

3301.0 

.000303 

.0886 

.2562 

60 

28.1 

1057.4 

1085.5 

1207.5 

.000828 

.2562 

.5056 

80 

48.1 

1046.6 

1094.7 

635.4 

.001573 

.5056 

1 

101.8 

69.8 

1034.6 

1104.4 

333.00 

.00300 

1 

2 

126.1 

94.1 

1021.4 

1115.5 

173.30 

.00577 

2 

3 

141.5 

109.5 

1012.3 

1121.8 

118.50 

.00845 

3 

4 

153.0 

120.9 

1005.6 

1126.5 

90.50 

.01106 

4 

5 

162.3 

130.2 

1000.2 

1130.4 

73.33 

.01364 

5 

6 

170.1 

138.0 

995.7 

1133.7 

61.89 

.01616 

6 

7 

173.8 

144.8 

991.7 

1136.5 

53.58 

.01867 

7 

8 

182.9 

150.8 

988.1 

1138.9 

47.27 

.02115 

8 

PROPERTIES  OF  STEAM 


43 


PROPERTIES  OF  SATURATED  STEAM  —  Continued 

ENGLISH   UNITS 


Abs.  Pressure 
Pounds  per 
Sq.  In. 

1 

Temperature 
Degrees  F. 

Heat  of  the 
Liquid 

Latent  Heat 
of  Evapora- 
tion 

—  S 

J-s 

* 

Density 
Pounds  per 
Cu.  Ft. 

Ahs.  Pressure  j 
Pounds  per 
Sq.  In.  | 

P 

t 

h- 

I 

H 

V 

3 

P 

9 

188.3 

156.3 

984.8 

1141.1 

42.36 

.023*61 

9 

10 

193.2 

161.2 

981.8 

1143.0 

38.38 

.02606 

10 

11 

197.7 

165.8 

979.0 

1144.8 

35.10 

.02849 

11 

12 

202.0 

170.0 

976.4 

1146.4 

32.38 

.03089 

12 

13 

205.9 

173.9 

974.0 

1147.9 

30.04 

.03329 

13 

14 

209.6 

177.6 

971.7 

1149.3 

28.02 

.03568 

14 

14.7 

212.0 

180.1 

970.4 

1150.4 

26.79 

.03733 

14.7 

15 

213.0 

181.1 

969.5 

1150.6 

26.27 

.03806 

15 

16 

216.3 

184.5 

967.4 

1151.9 

24.77 

.04042 

16 

17 

219.4 

187.7 

965.4 

1153.1 

23.38 

.04277 

17 

18 

222.4 

190.6 

963.5 

1154.1 

22.16 

.04512 

18 

19 

225.2 

193.5 

961.6 

1155.1 

21.07 

.04746 

19 

20 

228.0 

196.2 

959.8 

1156.0 

20.08 

.04980 

20 

21 

230.6 

198.9 

958.0 

1156.9 

19.18 

.05213 

21 

22 

233.1 

201.4 

956.4 

1157.8 

18.37 

.05445 

22 

23 

235.5 

203.9 

954.8 

1158.7 

17.62 

.05676 

23 

24 

237.8 

206.2 

953.2 

1159.4 

16.93 

.05907 

24 

25 

240.1 

208.5 

951.7 

1160.2 

16.30 

.0614 

25 

26 

242.2 

210.7 

950.3 

1161.0 

15.71 

.0636 

26 

27 

244.4 

212.8 

948.9 

1161.7 

15.18 

.0659 

27 

28 

246.4 

214.9 

947.5 

1162.4 

14.67 

.0682 

28 

29 

248.4 

217.0 

946.1 

1163.1 

14.19 

.0705 

29 

30 

250.3 

218.9 

944.8 

1163.7 

13.74 

.0728 

30 

31 

252.2 

220.8 

943.5 

1164.3 

13.32 

.0751 

31 

32 

254.1 

222.7 

942.2 

1164.9 

12.93 

.0773 

32 

33 

255.8 

224.5 

941.0 

1165.5 

12.57 

.0795 

33 

34 

257.6 

226.3 

939.8 

1166.1 

12.22 

.0818 

34 

35 

259.3 

228.0 

938.6 

1166.6 

11.89 

.0841 

35 

36 

261.0 

229.7 

937.4 

1167.1 

11.58 

.0863 

36 

37 

262.6 

231.4 

936.3 

1167.7 

11.29 

.0886 

37 

38 

264.2 

233.0 

935.2 

1168.2 

11.01 

.0908 

38 

39 

265.8 

234.6 

934.1 

1168.7 

10.74 

.0931 

39 

40 

267.3 

236.2 

933.0 

1169.2 

10.49 

.0953 

40 

41 

268.7 

237.7 

931.9 

1169.6 

10.25 

.0976 

41 

42 

270.2 

239.2 

930.9 

1170.1 

10.02 

.0998 

42 

43 

271.7 

240.6 

929.9 

1170.5 

9.80 

.1020 

43 

44 

273.1 

242.1 

928.9 

1171.0 

9.59 

.1043 

44 

45 

274.5 

243.5 

927.9 

1171.4 

9.39 

.1065 

45 

46 

275.8 

244.9 

926.9 

1171.8 

9.20 

.1087 

46 

44 


HEAT  ENGINES 


PROPERTIES  OF  SATURATED  STEAM  —  Continued 
ENGLISH  'UNITS 


Abs.  Pressure 
Pounds  per 
Sq.  In. 

Temperature 
Degrees  F. 

Heat  of  the 
Liquid 

Latent  Heat 
of  Evapora- 
tion 

Total  Heat 
of  Steam 

II^S 

cc  g 

Density 
Pounds  per 
Cu.  Ft. 

' 

Abs.  Pressure 
Pounds  per 
Sq.  In. 

P 

t 

h 

L 

H 

V 

1 
I 

P 

47 

277.2 

246.2 

926.0 

1172.2 

9.02 

.1109 

47 

48 

278.5 

247.6 

925.0 

1172.6 

8.84 

.1131 

48 

49 

279.8 

248.9 

924.1 

1173.0 

8.67 

.1153 

49 

50 

281.0 

250.2 

923.2 

1173.4 

8.51 

.1175 

50 

51 

282.3 

251.5 

922.3 

1173.8 

8.35 

.1197 

51 

52 

283.5 

252.8 

921.4 

1174.2 

8.20 

.1219 

52 

53 

284.7 

254.0 

920.5 

1174.5 

8.05 

.1241 

53 

54 

285.9 

255.2 

919.6 

1174.8 

7.91 

.1263 

54 

55 

287.1 

256.4 

918.7 

1175.1 

7.78 

.1285 

55 

56 

288.2 

257.6 

917.9 

1175.5 

7.65 

.1307 

56 

57 

289.4 

258.8 

917.1 

1175.9 

7.52 

.1329 

57 

58 

290.5 

259.9 

916.2 

1176.1 

7.40 

.1351 

58 

59 

291.6 

261.1 

915.4 

1176.5 

7.28 

.1373 

59 

60 

292.7 

262.2 

914.6 

1176.8 

7.17 

.1394 

60 

61 

293.8 

263.3 

913.8 

1177.1 

7.06 

.1416 

61 

62 

294.9 

264.4 

913.0 

1177.4 

6.95 

.1438 

62 

63 

295.9 

265.5 

912.2 

1177.7 

6.85 

.1460 

,63 

64 

297.0 

266.5 

911.5 

1178.0 

6.75 

.1482 

64 

65 

298.0 

267.6 

910.7 

1178.3 

6.65 

.1503 

65 

66 

299.0 

268.6 

910.0 

1178.6 

6.56 

.1525 

66 

67 

300.0 

269.7 

909.2 

1178.9 

6.47 

.1547 

67 

68 

301.0 

270.7 

908.4 

1179.1 

6.38 

.1569 

68 

69 

302.0 

271.7 

907.7 

1179.4 

6.29 

.1591 

69 

70 

302.9 

272.7 

906.9 

1179.6 

6.20 

.1612 

70 

71 

303.9 

273.7 

906.2 

1179.9 

6.12 

.1634 

71 

72 

304.8 

274.6 

905.5 

1180.1 

6.04 

.1656 

72 

73 

305.8 

275.6 

904.8 

1180.4 

5.96 

.1678 

73 

74 

306.7 

276.6 

904.1 

1180.7 

5.89 

.1699 

74 

75 

307.6 

277.5 

903.4 

1180.9 

5.81 

.1721 

75 

76 

308.5 

278.5 

902.7 

1181.2 

5.74 

.1743 

76 

77 

309.4 

279.4 

902.1 

1181.5 

5.67 

.1764 

77 

78 

310.3 

280.3 

901.4 

1181.7 

5.60 

.1786 

78 

79 

311.2 

281.2 

900.7 

1181.9 

5.54 

.1808 

79 

80 

312.0 

282.1 

900.1 

1182.2 

5.47 

.1829 

80 

81 

312.9 

283.0 

899.4 

1182.4 

5.41 

.1851 

81 

82 

313.8 

283.8 

898.8 

1182.6 

5.34 

.1873 

82 

83 

314.6 

284.7 

898.1 

1182.8 

5.28 

.1894 

83 

84 

315.4 

285.6 

897.5 

1183.1 

5.22 

.1915 

84 

85 

316.3 

286.4 

896.9 

1183.3 

5.16 

,1937 

85 

PROPERTIES  OF  STEAM 


45 


PROPERTIES  OF  SATURATED  STEAM  —  Continued 

ENGLISH    UNITS 


Abs.  Pressure 
Pounds  per 
Sq.  In. 

Temperature 
Degrees  F. 

Heat  of  the 
Liquid 

Latent  Heat 
of  Evapora- 
tion 

"cS  d 
«l 

is 

s° 

!i!h 

g.3£g 

ftSa* 

0 

Density 
Pounds  per 
Cu.  Ft. 

Abs.  Pressure 
Pounds  per 
Sq.  In. 

P 

t 

h 

L 

H 

• 

i 

V 

P 

86 

317.1 

287.3 

896.2 

1183.5 

5.10 

.1959 

86 

87 

317.9 

288.1 

895.6 

1183.7 

5.05 

.1980 

87 

88 

318.7 

288.9 

895.0 

1183.9 

5.00 

.2002 

88 

89 

319.5 

289.8 

894.3 

1184.1 

4.94 

.2024 

89 

90 

320.3 

290.6 

893.7 

1184.3 

4.89 

.2045 

90 

91 

321.1 

291.4 

893.1 

1184.5 

4.84 

.2066 

91 

92 

321.8 

292.2 

892.5 

1184.7 

4.79 

.2088 

92 

93 

322.6 

293.0 

891.9 

1184.9 

4.74 

.2110 

93 

94 

323.4 

293.8 

891.3 

1185.1 

4.69 

.2131 

94 

95 

324.1 

294.5 

890.7 

1185.2 

4.65 

.2152 

95 

96 

324.9 

295.3 

890.1 

1185.4 

4.60 

.2173 

96 

97 

325.6 

296.1 

889.5 

1185.6 

4.56 

.2194 

97 

98 

326.4 

296.8 

889.0 

1185.8 

4.51 

.2215 

98 

99 

327.1 

297.6 

888.4 

1186.0 

4.47 

.2237 

99 

100 

327.8 

298.4 

887.8 

1186.2 

4.430 

.2257 

100 

101 

328.6 

299.1 

887.2 

1186.3 

4.389 

.2278 

101 

102 

329.3 

299.8 

886.7 

1186.5 

4.349 

.2299 

102 

103 

330.0 

300.6 

886.1 

1186.7 

4.309 

.2321 

103 

101 

330.7 

301.3 

885.6 

1186.9 

4.270 

.2342 

104 

105 

331.4 

302.0 

885.0 

1187.0 

4.231 

.2364 

105 

106 

332.0 

302.7 

884.5 

1187.2 

4.193 

.2385 

106 

107 

332.7 

303.4 

883.9 

1187.3 

4.156 

.2407 

107 

108 

333.4 

304.1 

883.4 

1187.5 

4.119 

.2428 

108 

109 

334.1 

304.8 

882.8 

1187.6 

4.082 

.2450 

109 

110 

334.8 

305.5 

882.3 

1187.8 

4.047 

.2472 

110 

111 

335.4 

306.2 

881.8 

1188.0 

4.012 

.2493 

111 

112 

336.1 

306.9 

881.2 

1188.1 

3.977 

.2514 

112 

113 

336.8 

307.6 

880.7 

1188.3 

3.944 

.2535 

113 

114 

337.4 

308.3 

880.2 

1188.5 

3.911 

.2557 

114 

114.7 

337.9 

308.8 

879.8 

1188.6 

3.888 

.2572 

114.7 

115 

338.1 

309.0 

87§.7 

1188.7 

3878 

.2578 

115 

116 

338.7 

309.6 

879.2 

1188.8 

3.846 

.2600 

116 

117 

339.4 

310.3 

878.7 

1189.0 

3.815 

.2621 

117 

118 

340.0 

311.0 

878.2 

1189.2 

3.784 

.2642 

118 

119 

340.6 

311.7 

877.6 

1189.3 

3.754 

.2663 

119 

120 

341.3 

312.3 

877.1 

1189.4 

3.725 

.2684 

120 

121 

341.9 

313.0 

876.6 

1189.6 

3.696 

.2706 

121 

122 

342.5 

313.6 

876.1 

1189.7 

3.667 

.2727 

122 

123 

343.2 

314.3 

875.6 

1189.9 

3.638 

.2749 

123 

46 


HEAT  ENGINES 


PROPERTIES  OF  SATURATED  STEAM  —  Continued 

ENGLISH    UNITS 


Abs.  Pressure 
Pounds  per 
Sq.  In. 

Temperature 
Degrees  F. 

0> 
J3 
•^"O 

n 

r 

Latent  Heat 
of  Evapora- 
tion 

Total  Heat 
of  Steam 

B 

o 

Density 
Pounds  per 
Cu.  Ft. 

Abs.  Pressure 
Pounds  per 
Sq.  In. 

P 

t 

h 

L 

H 

V 

i 

V 

P 

124 

343.8 

314.9 

875.1 

1190.0 

3.610 

.2770 

124 

125 

344.4 

315.5 

874.6 

1190.1 

3.582 

.2792 

125 

126 

345.0 

316.2 

874.1 

1190.3 

3.555 

.2813 

126 

127 

345.6 

316.8 

873.7 

1190.5 

3.529 

.2834 

127 

128 

346.2 

317.4 

873.2 

1190.6 

3.5^3 

.2855 

128 

129 

346.8 

318.0 

872.7 

1190.7 

3.477 

.2876 

129 

130 

347.4 

318.6 

872.2 

1190.8 

3.452 

.2897 

130 

131 

348.0 

319.3 

871.7 

1191.0 

3.427 

.2918 

131 

132 

348.5 

319.9 

871.2 

1191.1 

3.402 

.2939 

132 

133 

349.1 

320.5 

870.8 

1191.3 

3.378 

.2960 

133 

134 

349.7 

321.0 

870.4 

1191.4 

3.354 

.2981 

134 

135 

350.3 

321.6 

869.9 

1191.5 

3.331 

.3002 

135 

136 

350.8 

322.2 

869.4 

1191.6 

3.308 

.3023 

136 

137 

351.4 

322.8 

868.9 

1191.7 

3.285 

.3044 

137 

138 

352.0 

323.4 

868.4 

1191.8 

3.263 

.3065 

138 

139 

352.5 

324.0 

868.0 

1192.0 

3.241 

.3086 

139 

140 

353.1 

324.5 

867.6 

1192.1 

3.219 

.3107 

140 

141 

353.6 

325.1 

867.1 

1192.2 

3.198 

.3128 

141 

142 

354.2 

325.7 

866.6 

1192.3 

3.176 

.3149 

142 

143 

354.7 

326.3 

866.2 

1192.5 

3.155 

.3170 

143 

144 

355.3 

326.8 

865.8 

1192.6 

3.134 

.3191 

144 

145 

355.8 

327.4 

865.3 

1192.7 

3.113 

.3212 

145 

146 

356.3 

327.9 

864.9 

1192.8 

3.093 

.3233 

146 

147 

356.9 

328.5 

864.4 

1192.9 

3.073 

.3254 

147 

148 

357.4 

329.0 

864.0 

1193.0 

3.053 

.3275 

148 

149 

357.9 

329.6 

863.5 

1193.1 

3.033 

.3297 

149 

150 

358.5 

330.1 

863.1 

1193.2 

3.013 

.3319 

150 

152 

359.5 

331.2 

862.3 

1193.5 

2.975 

.3361 

152 

154 

360.5 

332.3 

861.4 

1193.7 

2.939 

.3403 

154 

156 

361.6 

333.4 

860.5 

1193.9 

2.903 

.3445 

156 

158 

362.6 

334.4 

859.7 

1194.1 

2.868 

.3487 

158 

160 

363.6 

335.5 

858.8 

1194.3 

2.834 

.3529 

160 

162 

364.6 

336.6 

858.0 

1194.6 

2.801 

.3570 

162 

164 

365.6 

337.6 

857.2 

1194.8 

2.768 

.3613 

164 

166 

366.5 

338.6 

856.4 

1195.0 

2.736 

.3655 

166 

168 

367.5 

339.6 

855.5 

1195.1 

2.705 

.3697 

168 

170 

368.5 

340.6 

854.7 

1195.3 

2.674 

.3739 

170 

172 

369.4 

341.6 

853.9 

1195.5 

2.644 

.3782 

172 

174 

370.4 

342.5 

853.1 

1195.6 

2.615 

.3824 

174 

176 

371.3 

343.5 

852.3 

1195.8 

2.587 

.3865 

176 

PROPERTIES  OF  STEAM 


47 


PROPERTIES  OF  SATURATED  STEAM  —  Concluded 

ENGLISH   UNITS 


Abs.  Pressure 
Pounds  per 
Sq.  In. 

Temperature 
Degrees  F. 

Heat  of  the 
Liquid 

Latent  Heat 
of  Evapoi-a- 
tion 

Total  Heat 
of  Steam 

nh 
li*| 

£>3*< 

o 

Density 
Pounds  per 
Cu.  Ft. 

Abs.  Pressure 
Pounds  per 
Sq.  In. 

P 

t 

h 

L 

H 

V 

1  . 

P 

178 

372.2 

344.5 

851.5 

1196.0 

2.560 

.3907 

178 

180 

373.1 

345.4 

850.8 

1196.2 

2.532 

.3949 

180 

182 

374.0 

346.4 

850.0 

1196.4 

2.506 

.3990 

182 

184 

374.9 

347.4 

849.3 

1196.7 

2.480 

.4032 

184 

186 

375.8 

348.3 

848.5 

1196.8 

2.455 

.4074 

186 

188 

376.7 

349.2 

847.7 

1196.9 

2.430 

.4115 

188 

190 

377.6 

350.1 

847.0 

1197.1 

2.406 

.4157 

190 

192 

378.5 

351.0 

846.2 

1197.2 

2.381 

.4200 

192 

194 

379.3 

351.9 

845.5 

1197.4 

2.358 

.4242 

194 

196 

380.2 

352.8 

844.8 

1197.6 

2.335 

.4284 

196 

198 

381.0 

353.7 

844.0 

1197.7 

2.312 

.4326 

198 

200 

381.9 

354.6 

843.3 

1197.9 

2.289 

.4370 

200 

202 

382.7 

355.5 

842.6 

1198.1 

2.268 

.4411 

202 

204 

383.5 

356.4 

841.9 

1198.3 

2.246 

.4452 

204 

206 

384.4 

357.2 

841.2 

1198.4 

2.226 

.4493 

206 

208 

385.2 

358.1 

840.5 

1198.6 

2.206 

.4534 

208 

210 

386.0 

358.9 

839.8 

1198.7 

2.186 

.4575 

210 

212 

386.8 

359.8 

839.1 

1198.9 

2.166 

.4618 

212 

214 

387.6 

360.6 

838.4 

1199.0 

2.147 

.4660 

214 

216 

388.4 

361.4 

8%7.7 

1199.1 

2.127 

.4700 

216 

218 

389.1 

362.3 

837.0 

1199.3 

2.108 

.4744 

218 

220 

389.9 

363.1 

836.4 

1199.5 

2.090 

.4787 

220 

222 

390.7 

363.9 

835.7 

1199.6 

2.072 

.4829 

222 

224 

391.5 

364.7 

835.0 

1199.7 

2.054 

.4870 

224 

226 

392.2 

365.5 

834.3 

1199.8 

2.037 

.4910 

226 

228 

393.0 

366.3 

833.7 

1200.0 

2.020 

.4950 

228 

230 

393.8 

367.1 

833.0 

1200.1 

2.003 

.4992 

230 

232 

394.5 

367.9 

832.3 

1200.2 

.987 

.503 

232 

234 

395.2 

368.6 

831.7 

1200.3 

.970 

.507 

234 

236 

396.0 

369.4 

831.0 

1200.4 

.954 

.511 

236 

238 

396.7 

370.2 

830.4 

1200.6 

.938 

.516 

238 

240 

397.4 

371.0 

829.8 

1200.8 

.923 

.520 

240 

242 

398.2 

371.7 

829.2 

1200.9 

.907 

.524 

242 

244 

398.9 

372.5 

828.5 

1201.0 

.892 

.528 

244 

246 

399.6 

373.3 

827.8 

1201.1 

.877 

.532 

246 

248 

400.3 

374.0 

827.2 

1201.2 

.862 

.537 

248 

250 

401.1 

374.7 

826.6 

1201.3 

1.848 

.541 

250  _ 

275 

409.6 

383.7 

819.0 

1202.7 

1.684 

.594 

275  ~ 

300 

417.5 

392.0 

811.8 

1203.8 

1.547 

.647 

300 

350 

431.9 

407.4 

798.5 

1205.9 

1.330 

.750 

350 

CHAPTER  IV 
CALORIMETERS  AND  MECHANICAL  MIXTURES 

36.  Calorimeters. — As  we  have  already  seen,  steam  may  be 
either  wet,  dry  and  saturated,  or  superheated.     By  "quality"  of 
steam  is  meant  the  per  cent,  of  dry  and  saturated  steam  in  the 
sample. 

The  percent  of  moisture  in  the  steam  is  found  by  subtracting 
the  quality  from  100  per  cent. 

The  quality  is  determined  by  means  of  a  Calorimeter.  There 
are  two  classes  of  these  instruments  in  general  use  at  the  present 
time,  the  Separating  Calorimeter  and  the  Throttling  Calor- 
imeter. In  each  of  these  classes  there  are  several  types  or  makes, 
but  it  will  suffice  to  describe  only  one  or  two  of  each. 

As  will  be  seen  in  Paragraph  39,  the  American  Society  of 
Mechanical  Engineers  recommend  the  use  of  a  sampling  nozzle, 
or  calorimeter  nipple,  in  connection  with  the  calorimeter.  This 
nipple  is  a  piece  of  pipe  extending  nearly  across  the  steam  main, 
as  shown  in  Fig.  9,  with  a  cap  on  the  end  and  a  series  of  J-inch 
holes  along  and  around  its  cylindrical  surface.  As  the  steam  to 
be  tested  must  enter  the  calorimeter  through  this  nipple,  a  fair 
sample  of  the  steam  is  insured.  The  sampling  nozzle  should  be 
inserted  in  the  steam  main  at  a  point  where  the  entrained  mois- 
ture is  likely  to  be  most  thoroughly  mixed. 

37.  Separating  Calorimeters. — The  weight  of  the  dry  steam 
that  will  pass  through  a  given  size  of  orifice  in  a  given  time 
depends  upon  the  pressure  on  the  two  sides  of  the  orifice.     If 
A  is  the  area  of  the  orifice  in  square  inches,  P  the  absolute 
pressure  in  pounds  per  square  inch,  and  W  the  pounds  of  steam 
passing  through   the  orifice  into  the  atmosphere  per  second, 
then 

PA 
W=-7Q  (Napier's  Rule).  (1) 

From  Napier's  Rule  the  weight  of  steam  flowing  through  an 
orifice  of  known  area  is  proportional  to  the  absolute  steam  pres- 
sure. This  law  holds  true  until  the  lower  pressure  equals  or 
exceeds  .6  of  the  higher  pressure. 

48 


CALORIMETERS  AND  MECHANICAL  MIXTURES      49 


The  amount  of  steam  flowing  through  any  orifice  may,  there- 
fore, be  determined.  Professor  R.  C.  Carpenter  has  a  calorim- 
eter based  upon  this  principle.  Wet  steam  enters  the  calorim- 
eter, Fig.  7,through  the  pipe  6,,and  is  projected  against  the  cup 
14.  The  steam  and  water  are  then  turned  through  an  angle 
of  180°,  which  causes  the  water  to  be  thrown  outward  by  cen- 
trifugal force  through  the  meshes  in  the  cup  into  the  inner  cham- 
ber 3.  Causing  the  steam  to  strike  the  cup,  instead  of  flowing 
directly  into  the  chamber  3,  prevents  any  moisture  already 
thrown  out  being  picked  up  again 
and  carried  on.  The  steam  after 
leaving  the  cup  passes  upward  and 
enters  the  top  of  the  outer  chamber 
7.  It  then  flows  down  around  the 
inner  chamber  in  the  annular  space 
4,  and  is  discharged  through  the  ori- 
fice 8.  The  area  of  this  orifice,  which 
is  known,  is  so  small  that  there  is  no 
loss  in  pressure  of  the  steam  as  it  flows 
through  the  calorimeter.  The  pres- 
sure in  the  two  chambers  being  the 
same,  the  temperature  is  the  same, 
and  there  is  no  loss  of  heat  from  the 
inner  chamber  by  radiation.  The 
gage  glass  12,  connected  with  the  in- 
ner chamber,  is  graduated  in  hun- 
dredths  of  pounds,  so  that  the  weight 
of  moisture  separated  from  the  steam 
-can  be  read  directty.  The  gage  9  is 
so  calibrated  as  to  read  directly  the 


FIG.  7. — Carpenter's  improved 
separating  calorimeter. 


number  of  pounds  flowing  through  the  orifice  8  in  a  given  time 
(generally  ten  minutes).  These  readings  are  not  proportional 
to  the  pressure  readings  on  the  gage,  which  has  two  scales,  for 
the  latter  readings  are  proportional  to  the  pressures  above  the 
atmosphere  and  not  to  the  absolute  pressures.  The  accuracy 
of  the  results  obtained  by  \ising  the  gage  may  be  checked  at 
any  time  by  condensing  and  weighing  the  discharge  from  ori- 
fice 8  for  a  given  period  of  time. 

If  now  we  call  w  the  weight  of  dry  steam  discharged  from 
the  orifice  8  in  any  given  period  of  time,  W  the  weight  of  mois- 


50 


HEAT  ENGINES 


ture  collected  in  3  in  the  same  period  of  time,  and  q  the  quality 
of  the  steam,  then 


w  may  be  obtained  either  from  the  reading  of  the  gage  9,  or 
by  actually  weighing  the  steam,  and  W  is  found  by  taking 
the  difference  between  the  readings  on  scale  12  at  the  beginning 
and  end  of  the  test. 

38.  Throttling  Calorimeter.  —  This  form  of   calorimeter   was 
invented  by  Prof.  C.  H.  Peabody,  and  is  the  form  recommended 


FIG.  8. — Carpenter's  throttling  calorimeter. 


by  the  A.S.M.E.  Committee  on  Standards  (see  paragraph  39 
below).  It  is  the  most  accurate  form  of  calorimeter  where  it 
can  be  used,  but  is  unsuitable  for  use  in  determining  the  quality 
of  the  steam  if  the  steam  contains  over  3  or  4  per  cent,  of  mois- 
ture, or  if  the  temperature  of  the  lower  thermometer  is  below 
225°,  which  will  be  the  case  if  the  steam  is  at  a  very  low  pres- 
sure (below  5  or  6  Ibs.  gage). 

The  principle  of  its  operation  is  as  follows:  a  pound  of  saturated 
steam  at  a  high  pressure  contains  more  heat  than  a  pound 
of  saturated  steam  at  a  lower  pressure.  If  steam  at  a  high 
pressure  pass  through  an  orifice  into  a  space  at  a  lower  pressure 
without  doing  any  external  work,  some  of  this  heat  must  be  given 
up,  and  as  the  only  object  that  can  absorb  heat  is  the  steam  itself, 
it  takes  up  this  heat.  If  this  steam  contained  some  moisture 
at  the  higher  pressure,  part  of  the  heat  liberated  when  the  pres- 


CALORIMETERS  AND  MECHANICAL  MIXTURES      51 

sure  is  lowered  will  go  to  evaporating  this  moisture,  and  the 
excess  will  go  to  superheating  the  steam. 

Let  q  =  the  quality  of  the  steam. 

ti  =  the  temperature  of  the  wet  steam  before  passing  through 

trie  orifice. 

Pi  =  the  absolute  pressure  of  the  wet  steam  in  the  main. 
fa  =  the  temperature  corresponding  to  the  absolute  pressure 

on  the  low-pressure  side  of  the  orifice. 

i  sup-  =  the  temperature  of  the  steam  as  shown  by  the  thermom- 
eter on  the  low-pressure  side  of  the  orifice. 
hi  and  Z/i  =  heat  of  liquid  and  latent  heat  corresponding  to  the 

temperature   fa,  or  the  absolute  pressure  p\. 
h2  and  L2  =  heat  of  liquid  and  latent  heat  corresponding  to 
the  temperature  fa. 

The  heat  contained  in  1  Ib.  of  the  mixture  of  steam  and 
water  at  temperature  fa,  or  pressure  pi,  would  be 

hi  +  qLi. 

The  heat  contained  in  1  Ib.  of  the  steam  on  the  low-pressure 
side  of  the  orifice  after  expansion  would  be 

A2  +  Lz  +  cp(tsup.  -  fa)  =  H2  +  cp(tsup.  -  fa), 

where  cp  is  the  specific  heat  of  superheated  steam.  But  since  the 
heat  in  a  pound  of  the  substance  must  be  the  same  on  one  side  of 
the  orifice  as  it  is  on  the  other, 

hi  +  qLi  =  H2  +  cp  (tsup.  -  fa).  (3) 

Solving  for  #, 

H2  +  cp  (tsup.  -  fa)  -  hi 
q=  Li 

The  percentage  of  moisture  equals  1  —  q.  (5) 

Ordinarily  fa  is  found  from  the  tables  by  looking  up  the 
temperature  corresponding  to  the  absolute  pressure  in  the  calo- 
rimeter, i.e.,  the  sum  of  the  atmospheric  pressure  and  the  pres- 
sure shown  by  the  manometer.  This  practice,  however,  is  not 
permitted  by  the  A.S.M.E.  rules  for  finding  the  quality  of 
steam,  since  tsup.  is  taken  with  a  thermometer  that  has  part  of 


52 


HEAT  ENGINES 


its  stem  exposed,  and  is  thus  subject  to  radiation,*  nor  does  it 
take  account  of  the  radiation  from  the  calorimeter  itself,  which 
may  be  considerable  even  though  well  covered.  Therefore  for 
accurate  work  it  is  necessary  that  we  take  a  "normal  reading" 
of  the  thermometer,  as  described  in  paragraph  39  to. correct  for 
these  errors. 

The  calorimeter  shown  in  Fig.  9  differs  from-  the  one  shown 
in  Fig.  8  in  that  the  temperature  of  the  steam  being  admitted  to 
the  calorimeter  is  observed  instead  of  the  pressure.  In  other 


f 


FIG.  9. — Barms'  throttling  calorimeter. 

words,  hi  and  LI  correspond  to  the  temperature  ti  rather  than  to 
the  absolute  pressure  p\.  Another  difference  is  that  in  the 
Barrus  calorimeter,  the  exhaust  is  made  very  free  and  the  pres- 
sure, pz,  on  the  lower  side  of  the  orifice  is  assumed  to  be  atmos- 
pheric. A  long  exhaust  pipe  will  cause  a  back  pressure  in  the 
calorimeter  where  we  have  assumed  the  pressure  to  be 
atmospheric. 

In  case  the  atmospheric  pressure  is  not  known,   it  can  be 

*When  a  considerable  portion  of  the  mercury  column  of  a  thermometer 
measuring  high  temperatures  is  exposed  to  the  air,  a  correction  K  must  be 
added  to  the  readings  to  obtain  the  true  temperature. 
Let  t  =  the  observed  reading  of  the  thermometer. 

t'  =  the  temperature  of  the  air  surrounding  the  exposed  stem  of  the 

thermometer. 

D  =  number  of  degrees  on  the  scale  from  the  surface  of  the  liquid  in 
the  thermometer  cup  to  the  upper  end  of  the  mercury  column. 
Then  K  =  .000088  D  (t  —  t') ,  in  Fahrenheit  degrees. 


CALORIMETERS  AND  MECHANICAL  MIXTURES      53 

assumed  as  14.7  Ibs.  per  square  inch.  If  the  barometer  reading 
is  given,  however,  it  should  always  be  used.  This  reading,  as 
well  as  that  of  the  manometer  giving  the  pressure  in  the  calo- 
rimeter, will  be  given  in  inches  of  mercury.  To  change  this  to 
pounds  per  square  inch,  multiply  the  inches  of  mercury  by  .491. 
39.  Quality  of  Steam. — The  following  are  the  standard  rules 
for  finding  the  quality  of  steam  as  adopted  by  the  A.S.M.E., 
and  published  in  the  Transactions  of  that  society,  Vol.  21,  p.  43, 
and  Vol.  24,  p.  740: 

"  The  percentage  of  moisture  in  the  steam  should  be  determined  by  the 
use  of  either  a  throttling  or  a  separating  steam  calorimeter.  The  sam- 
pling nozzle  should  be  placed  in  the  vertical  steam  pipe  rising  from  the 
boiler.  It  should  be  made  of  |-in.  pipe,  and  should  extend  across  the 
diameter  of  the  steam  pipe  to  within  half  an  inch  of  the  opposite  side, 
being  closed  at  the  end  and  perforated  with  not  less  than  twenty  f-in. 
holes  equally  distributed  along  and  around  its  cylindrical  surface,  but 
none  of  these  holes  should  be  nearer  than  \  in.  to  the  inner  side  of  the 
steam  pipe.  The  calorimeter  and  the  pipe  leading  to  it  should  be  well 
covered  with  felting.  Whenever  the  indications  of  the  throttling  or 
separating  calorimeter  show  that  the  percentage  of  moisture  is  irregular, 
or  occasionally  in  excess  of  3  per  cent.,  the  results  should  be  checked 
by  a  steam  separator  placed  in  the  steam  pipe  as  close  to  the  boiler  as 
convenient,  with  a  calorimeter  in  the  steam  pipe  just  beyond  the  outlet 
from  the  separator.  The  drip  from  the  separator  should  be  caught  and 
weighed  and  the  percentage  of  moisture  computed  therefrom  added  to 
that  shown  by  the  calorimeter. 

"Superheating  should  be  determined  by  means  of  a 'thermometer 
placed  in  a  mercury-well  inserted  in  the  steam  pipe.  The  degree  of 
superheating  should  be  taken  as  the  difference  between  the  reading  of  the 
thermometer  for  superheated,  steam  and  the  readings  of  the  same 
thermometer  for  saturated  steam  at  the  same  pressure,  as  determined 
by  a  special  experiment,  and  not  by  reference  to  steam  tables." 

"If  it  is  necessary  to  attach  the  calorimeter  to  a  horizontal  section  of 
pipe,  and  it  is  important  to  determine  the  quantity  of  moisture  accu- 
rately, a  sampling  nozzle  should  be  used  which  has  no  perforations,  and 
which  passes  through  a  stuffing-box  applied  to  the  bottom  of  the  pipe  so 
that  it  can  be  adjusted  up  and  down,  and  thereby  draw  a  sample  at 
different  points  ranging  from  the  top  to  the  bottom. 

"To  determine  the  'normal  reading'  of  the  calorimeter,  the  instrument 
should  be  attached  to  a  horizontal  steam  pipe  in  such  a  way  that  the 
nozzle  projects  upward  to  near  the  top  of  the  pipe,  there  being  no  per- 
forations and  the  steam  entering  through  the  open  end.  The  test  should 
be  made  when  the  steam  in  the  pipe  is  in  a  quiescent  state,  and  when  the 


54  HEAT  ENGINES 

steam  pressure  is  constant.  If  the  steam  pressure  falls  during  the  time 
when  the  observations  are  being  made,  the  test  should  be  continued  long 
enough  to  obtain  the  effect  of  an  equivalent  rise  of  pressure.  When  the 
normal  reading  has  been  obtained,  the  constant  to  be  used  in  determining 
the  percentage  of  moisture  is  the  latent  heat  of  the  steam  at  the  observed 
pressure  divided  by  the  specific  heat  of  superheated  steam  at  atmospheric 
pressure,  which  is  forty-six  hundredths  (.46).  To  ascertain  this 
percentage,  divide  the  number  of  degrees  of  cooling  by  the  constant,  and 
multiply  by  100. 

"To  determine  the  quantity  of  steam  used  by  the  calorimeter  in  an 
instrument  where  the  steam  is  passed  through  an  orifice  under  a  given 
pressure,  it  is  usually  accurate  enough  to  calculate  the  quantity  from  the 
area  of  the  orifice  and  the  absolute  pressure.  If  it  is  desired  to  determine 
the  quantity  exactly,  a  steam  hose  may  be  attached  to  the  outlet  of  the 
calorimeter,  and  carried  to  a  barrel  of  water  placed  on  a  platform  scale. 
The  steam  is  condensed  for  a  certain  time,  and  its  weight  determined, 
and  thereby  the  quantity  discharged  per  hour." 

Example. — Steam  at  100  Ibs.  pressure  blows  through  a  throttling 
calorimeter.  The  temperature  of  the  lower  thermometer  is  275°  and 
the  manometer  reading  is  5.6  in.  of  mercury.  Barometer  reading 
29  in.  Find  the  quality  of  the  steam. 

Solution. — First  find  the  atmospheric  pressure  and  the  pressure  in  the 
calorimeter. 

Atmospheric   pressure  =  .491  X  29  =  14.25  Ibs. 
Pressure  in  calorimeter  =  .491  X  5.6  =  2.75  Ibs. 

Now  from  the  steam  tables  find  hi  and  LI  corresponding  to  the  pressure  in 
the  main,  114.25  Ibs.  absolute,  and  also  Hz  and  £2  corresponding  to  the 
pressure  in  the  calorimeter,  17  Ibs.  absolute. 
Then  from  equation  (4), 

Hz  +  cp  (tsup.  -  «2)  -  hi 


1153.1  +  .46  (275  -  219.4)  -  308.5 

RQ  a 

=  .988. 


880.1 
1153.1  +  .46  X  55.6  -  308.8        869.6 


880.1  880.1 

Answer:  —  98.8  per  cent. 

Example. — (a)  Find  the  quality  of  the  steam  in  the  preceding  problem 
as  shown  by  a  separating  calorimeter,  if  the  data  is  as  follows :  weight  of 
dry  steam  escaping  through  orifice,  4.5  Ibs.;  weight  of  moisture  col- 
lected, .05  Ibs. 

(6)  Find  the  diameter  of  the  orifice  if  the  length  of  the  run  is 
20  minutes. 


CALORIMETERS  AND  MECHANICAL  MIXTURES       55 

Solution. — (a)  From  equation  (2), 

w 
q  ~  w  +  W 

=  4^+^05   =4.55   =  -988' 

(6)  Find  weight  of  steam  flowing  through  orifice  per  second,  and  call 
it  w'.     Then 

45  45 

w'  =  6n"<7«n  =  — ^  =  -00375  Ibs . 
Zi\J  X  OU 

From  equation  (1), 

PA 

W  =    70 


70  X  .OOS^1     -26250 


P          1004-  .49O<29       114.25  gV   (. 

*r*  =  .0023  %V^*>       &     *  ^-r  Pr 

"--S"-000732  * 

r  =  .027 
d  =  .054 

Answer:  (a)  98.8  per  cent 
(6)  .054  in. 

PROBLEMS 

1.  Steam  at  100  Ibs.  pressure  passes  through  a  Barrus  calorimeter.  Tem- 
perature after  passing  through  orifice  is  246°.  What  is  the  quality  of  the 
steam?  Q^ni 

(2)  Steam  at  110  Ibalbliws  through  an  orifice  into  the  atmosphere.  The 
temperature  of  the  steam  after  passing  through  this  orifice  is  240°.  What 
per  cent,  of  moisture  is  in  the  original  steam? 

3.  One  pound  of  a  mixture  of  steam  and  water  containing  2  per  cent, 
moisture  at  150  Ibs.  absolute  pressure  expands  through  an  orifice  to  15  Ibs. 
absolute  pressure.     What  will  be  the  temperature  at  the  lower  pressure? 
^    4.  Steam  at  a  pressure  of  100  Ibs.  and  a  quality  of  98  per  cent,  blows 
through  an  orifice  to  15  Ibs.  absolute.     What  will  be  its  temperature? 
^  vJ  ^eam  a^  95  Ibs.  pressure  containing  2|  per  cent,  moisture  blows 
through  an  orifice  into  a  chamber  where  the  pressure  is  8.2  in.  of  mercury 
above  the  atmosphere.     What  is  the  temperature  of  the  steam  after  passing 
through  the  orifice?     Barometer,  29.8  in. 

»     6.  Find  the  quality  of  the  steam  if,  when  tested  with  a  separating  calor- 
imeter, 4.5  Ibs.  of  dry  steam  blow  through  the  orifice  while  1.5  Ibs.  of  mois- 


56  HEAT  ENGINES 

ture  are  separated  out.     If  the  run  is  thirty  minutes  long  and  the  steam 
pressure  is  100  Ibs.,  determine  the  diameter  of  the  orifice. 

7.  Steam  at  10  Ibs.  pressure  blows  through  a  separating  calorimeter. 
The  run  is  forty-five  minutes  long,  10.5  Ibs.  of  dry  steam  flow  through  the 
orifice  and  .5  Ibs.  of  moisture-are  collected.  Find  the  quality  of  the  steam 
and  the  area  of  the  orifice. 

40.  Mechanical  Mixtures. — Problems  involving  the  resulting 
temperature  and  final  condition  when  various  substances  are 
mixed  mechanically  are  often  met  with.  They  are  best  treated 
by  first  determining  the  heat  in  B.T.U.  that  would  be  available 
for  use  if  the  temperature  of  all  the  substances  were  brought  to 
32°  F.,  and  then  using  this  heat  (positive  or  negative)  to  raise 
(or  lower)  the  total  weight  of  mixture  to  its  final  temperature 
and  condition. 

Another  method  of  solving  mixture  problems  is  by  equating 
the  heat  absorbed  to  the  heat  rejected  and  letting  z. represent  the 
resulting  temperature.  It  is  often  difficult  to  decide  upon 
which  side  of  the  equation  a  material  should  be  placed.  In 
such  a  case  a  trial  calculation  should  be  made,  and  the  tempera- 
ture determined  by  this  trial  will  settle  this  question. 

In  the  mixture  of  substances  which  pass  through  a  change  of 
state  during  the  mixture  process,  it  is  almost  necessary  to  make 
a  trial  calculation.  Take,  for  example,  the  mixing  of  steam  with 
other  substances.  The  steam  may  all  be  condensed  and  the  re- 
sulting water  cooled  also;  the  steam  may  be  condensed  only;  or 
the  steam  may  be  only  partially  condensed.  The  equations  in 
each  case  would  be  different. 

If  one  pound  of  dry  saturated  steam  at  a  temperature  ti  is 
condensed  and  then  the  temperature  of  the  condensed  steam 
is  lowered  to  a  temperature  £2,  the  amount  of  heat  Hr  given  off 
would  be 

H'  =  Li  +  c(t,  -  tz).  (6) 

where  LI  is  the  latent  heat  corresponding  to  the  temperature  t\. 

If  the  steam  was  condensed  only,  the  heat  given  off  would  be 

H'  =  Li  (7) 

and  the  temperature  of  the  mixture  is  the  temperature  corre- 
sponding to  the  pressure. 

If  the  steam  is  only  partly  condensed,  let  q  equal  the  per  cent, 
of  steam  condensed.  Then 

H'  =  qLi  (8) 

and  the  temperature  of  the  mixture  is  the  temperature  corre- 
sponding to  the  pressure. 


CALORIMETERS  AND  MECHANICAL  MIXTURES      57 

The  general  laws  of  thermodynamics  do  not  apply  in  the 
case  of  mixtures  as  the  equations  become  discontinuous.  The 
general  expression  for  heat  absorbed  in  passing  from  a  solid  to 
a  gaseous  state  may  be  stated  as  follows: 

Let  Ci  be  the  specific  heat  in  the  solid,  c2  in  the  liquid  and  c3 
in  the  gaseous  state,  w  the  weight  of  the  substance,  t  the  initial 
temperature,  t\  the  temperature  of  the  melting  point,  t2  the 
temperature  of  the  boiling  point,  t3  the  final  temperature,  Hf 
heat  of  liquefaction,  and  L  heat  of  vaporization. 

Hr  =  w[d  (t,  -  t)  +  Hf  +  c2  (tz  -  O  +  L  +  c3  (h  -  fe)]    (9) 
TABLE  VIII.  SPECIFIC  HEATS  OF  LIQUIDS  AND  SOLIDS 


Substances 

Specific  heat,  c. 

Mercury  

.0333 

Alcohol 

.615 

Turpentine  
Wrought  iron 

.462 
.114 

Cast  iron  

.129 

Copper  
Ice 

.095 
504 

Spermaceti  
Sulphur 

.320 
.177 

Glass  
Graphite 

.187 
.200 

Latent  heat  of  fusion  of  ice  =  144  B.T.U. 

Example.— Find  the  final  temperature  and  condition  of  the  mixture 
after  mixing  10  Ibs.  of  ice  at  20°;  20  Ibs.  of  water  at  50°,  and  2  Ibs.  of 
steam  at  atmospheric  pressure.  Mixture  takes  place  at  the  pressure  of 
the  steam. 

Solution.—  First  Method 

Heat  to  raise  ice  to  32°  =  10  X  .5  (32  -  20)  60 

Heat  to  melt  ice  =  10  X  144  =  1440 

Total  heat  necessary  to  change  the  ice  to  water 

at  32°  =  1500  B.T.U. 

Heat  given  up  by  water  when  temperature  is 

lowered  to  32°  =  20  X  (50  -  32)  =  360 

Heat  in  steam  above  32°  (from  tables) 

=  2X1150.4  =2300.8 

Total  heat  given   up    in    lowering  water   and 

steam  to  32°  =  2660.8  B.T.U. 

Heat  available  for  use  =  2660.8  -  1500  =  1160.8  B.T.U. 


58  HEAT  ENGINES 

1160  8 
Degrees  this  heat  will  raise  the  mixture  =  36.3. 

oZ 

.'.  final  temperature  of  mixture  =  36.3  -f  32  =  68.3°  F. 
Ans.  32  Ibs.  water  at  68.3°  F. 


Second  Method 

Assume  that  the  steam  is  all  condensed  and  that  the  temperature  of  the 
mixture  is  t°.  Then  the  heat  necessary  to  raise  the  ice  to  the  melting 
point  equals 

10  X  .5(32  -  20). 

The  heat  necessary  to  melt  the  ice  equals  10  X  144;  the  heat  necessary 
to  raise  £he  melted  ice  to  the  temperature  of  the  mixture  equals 
10(£  —  32);  the  heat  necessary  to  raise  the  water  to  the  temperature  of 
the  mixture  equals  20  (t  —  50) ;  the  heat  given  up  by  the  steam  in  chang- 
ing to  water  at  the  temperature  of  the  boiling  point  equals  2  X  970.4, 
and  the  heat  given  up  by  the  condensed  steam  when  its  temperature  is 
lowered  to  the  temperature  of  the  mixture  equals  2  (212  —  t). 
Combining  the  preceding  parts  into  one  equation,  we  have 

10  X  .5  (32  -  20)  +  10  X  144  +  10  (t  -  32)  +  20  (t  -  50)  = 
2  X  970.4  +  2  (212  -  t) 

60 '+  1440  +  Wt  -  320  +  20*  -  1000  =  1940.8  +  424  -  2t 

32£  =  2184.8 

t  =  68.3° 

Since  t  is  less  than  the  temperature  of  the  boiling  point  corresponding 
to  the  pressure  at  which  the  mixture  takes  place,  all  the  steam  is  con- 
densed. 

Ans.  32  Ibs.  water  at  68.3°  F. 

Example. — Find  the  resulting  temperature  and  condition  after  mixing 
10  Ibs.  of  ice  at  20°,  20  Ibs.  of  water  at  50°,  40  Ibs.  of  air  at  82°,  and  20 
Ibs.  of  steam  at  100  Ibs.  pressure  and  containing  2  per  cent,  moisture. 
Mixture  takes  place  at  the  pressure  of  the  steam. 

Solution. —  First  Method 

10  X  .5(32  -  20)  =  60 

10  X  144  =  1440 

1500  B.T.U.  =  heat  to  raise  ice  to   water 
at  32°. 


CALORIMETERS  AND  MECHANICAL  MIXTURES      59 

20  X  (50  -  32)  360 

40  X  .2375(82  -  32)  475 

20(308.8  +.98  X  879.8)  =   23420 

24255  B.T.U.  =  heat    given    up    by    air, 

1500  water,  and  steam. 

22755  B.T.U.  =  heat  available. 

40  X  .2375(337.9  -  32)  =    2905  B.T.U.  =  heat  to  raise  air  to,337.9°. 

19850  B.T.U.  =  heat  available  to  raise  the 

water. 
50  X  308.8  =  15440  B.T.U.  =  heat    to    raise    water    to 

337.9°. 

4410  B.T.U.  =  heat  available  to  evapo- 
rate water. 
4410 

oTft  o  =  5.01  Iks.  steam, 
o/y .  o 

Ans.    40  Ibs.  air 

44  .99  Ibs.  water  >  at  337  .9°. 

5.01  Ibs  dry  saturated  steam  J 

Second  Method 

Assume  the  steam  to  be  all  condensed  and  let  the  temperature  of 
the  mixture  be  t°.  Equating  the  heat  gained  by  the  ice,  water  and  air, 
and  the  heat  lost  by  the  steam,  we  have 

10  X   .5(32  -  20)  +  10  X  144  +  10(«  -  32)  +  20(*  -  50)  +  40  X 
.2375(1  -  82)  =  20  X  .98  X  879.8  +  20(337.9  -  0 

60  +  1440  +  IQt  -  320  +  20*  -  1000  +  9.5Z  -  779  =   17250  + 
6758  -  20t 

59.5*  =  24670 
t  =  413. 6°  F. 

This  result  is  of  course  absurd,  as  the  temperature  of  the  mixture  cannot 
be  higher  than  the  temperature  of  the  boiling  point  corresponding  to  the 
pressure  at  which  the  mixture  takes  place.  Therefore  our  assumption 
that  all  the  steam  is  condensed  must  be  wrong,  and  we  know  that  part 
of  it  remains  in  the  form  of  steam,  and  hence  the  temperature  of  the  mix- 
ture is  equal  to  the  temperature  of  the  boiling  point  corresponding  to  the 
pressure  at  which  the  substances  are  mixed. 

Then  substituting  for  t  its  value,  and  letting  x  represent  the  number  of 
pounds  of  steam  condensed,  we  have 

10  X   .5(32  -  20)  +  10  X  144  +  10(337.9  -  32)  +  20(337.9  -  50) 

+  40  X  .2375(337.9  -  82)  =  879. 8* 


60  HEAT  ENGINES 

60  +  1440  +  3059  +  5758  +  2431  =  879. 8x 
879. 8  x  =  12748 

x  =  14.49  Ibs.  condensed. 
20  X  .98  =  19.6  Ibs.  =  original  weight  of  dry  steam. 

Ans.  40  Ibs.  air  ) 

10  +  20  +  (20  -  19.6)  +  14.49  =  44.89  Ibs.  water  [  at  337.9°. 
19.6  -  14.49  =  5.11  Ibs.  dry  saturated  steam 

The  difference  between  the  results  obtained  in  these  two  methods  of 
working  this  problem  is  due  to  the  fact  that  in  the  first  method  we  took 
account  of  the  variation  in  the  specific  heat  of  water  by  using  the  heat  of 
the  liquid,  h,  from  the  tables,  in  place  of  (t  —  32)  wherever  possible, 
while  in  the  second  method  we  assumed  this  specific  heat  to  be  constant 
and  equal  to  1. 

Example. — Find  the  resulting  temperature  and  condition  after  mixing 
10  Ibs.  of  ice  at  20°,  20  Ibs.  of  water  at  50°,  and  30  Ibs.  of  steam  at  100  Ibs. 
pressure  and  400°  temperature.  Mixture  takes  place  at  25  Ibs.  pressure. 

Solution.—  First  Method 

10  X  .5(32  -  20)  =      60 

10  X  144  =  1440 

1500  B.T.U.  =  heat  to  raise  ice  to  water  at 

32°. 

20  X  (50  -  32)  =      360 

30  X  .57(400  -  337.9)  =    1062 
30  X  1188.6  35658 

37080  B.T.U.  =  heat  given  up  by  water  and 

steam. 
1500 

35580  B.T.U.  =  heat  available. 
60  X  235.7  =  14142  B.T.U.  =  heat    to    raise    water    to 

266.8°. 
21438  B.T.U.  =  heat  available  to  evaporate 

water. 
21438 


22 . 97  Ibs.  steam 

'  at  266. 8°  F. 


933.3 

Ans.     37 . 03  Ibs.  water 

22 . 97  Ibs.  dry  saturated  steam 

Second  Method 

Assume  the  steam  to  be  all  condensed  and  let  the  temperature  of  the 
mixture  be  t°.     Then 


CALORIMETERS  AND  MECHANICAL  MIXTURES      61 

10  X   .5(32  -  20)  +  10  X  144  +  10(*  -  32)  +  20(J  -  50)  =  30  X 

.57(400  -  337.9)  +  30  X  879.8  +  30(337.9  -  0 
60  +  1440  +  10*  -  320  +  20£  -  1000  =  1062  +  26394  +  10137  - 

60  t  =  37413 
t  =  623.6° 

This  result  is,  of  course,  impossible  and  we  see  at  once  that  only  part 
of  the  steam  is  condensed,  and  that  the  temperature  of  the  mixture  must 
be  that  of  the  boiling  point  corresponding  to  the  pressure  at  which  the 
mixture  takes  place. 

This  problem  differs  from  the  previous  ones  in  that  the  pressure  of  the 
mixture  is  different  from  the  original  steam  pressure,  and  we  must  pro- 
ceed in  a  slightly  different  manner. 

Assume  for  the  moment  that  the  steam  has  all  been  condensed  and  that 
we  have  60  Ibs.  of  water  at  623 . 6°  F.  Then  assume  that  the  temperature 
of  the  water  is  dropped  to  the  temperature  of  the  boiling  point  (266.8°) 
corresponding  to  the  pressure  (25  Ibs.)  at  which  the  mixture  is  made. 
Each  pound  will  give  up,  approximately  (623 .6  -  266 . 8)  B.T.U.  This 
heat  can  then  be  used  to  re-evaporate  part  of  the  water.  Therefore  since 
the  latent  heat  corresponding  to  25  Ibs.  is  933 . 3,  we  have 

60(623,6-266.8)       60X356.8       21408 

:*)33.3  933.3        =  933.3  =  22. 94 Ibs.  re-evaporated. 

VY     ^ 

An*.  37.06  Ibs.  water  \ 

1    ^2.94  Ibs.  dry  saturated  steam  / 


PROBLEMS 

1.  Required  the  temperature  after  mixing  3  Ibs.  of  water  at  100°  F.,  10 
ll>s.  of  alcohol  at  40°  F.,  and  20  Ibs.  of  mercury  at  60°  F. 

2.  Required  the  temperature  and  condition  after  mixing  5  Ibs.  of  ice  at 
10°  F.  with  12  Ibs.  of  water  at  60°  F. 

3.  Required  the  temperature  and  condition  after  mixing  10  Ibs.  of  ice  at 
15°  F.  with  1  Ib.  of  steam  at  212°  F. 

4.  Required  the  temperature  and  condition  of  the  mixture  after  mixing 
5  Ibs.  of  steam  at  212°  F.  with  20  Ibs.  of  water  at  60°  F. 

One  pound  of  ice  at  32°  is  mixed  with  10  Ibs.  of  water  at  50° and  20  Ibs. 
of  steam  at  212°.  What  is  the  temperature  and  condition  of  the  resulting 
mixture? 

6.  Ten  pounds  of  steam  at  212°  are  mixed  with  50  Ibs.  of  water  at  60° 
and  2  Ibs.  of  ice  at  32°.     What  will  be  the  resulting  temperature  and  condi- 
tion of  the  mixture? 

7.  Ten  pounds  of  steam  at  atmospheric  pressure,  5  Ibs.  of  water  at  50° 
and  10  Ibs.  of  ice  at  32°  are  mixed  together,     (a)  What  will  be  the  resulting 

JV    *\temperature  of  the  mixture?     (6)  What  will  the  condition  of  the  mixture 
V  be?     (c)  If  the  steam  is  not  all  condensed,  determine  what  per  cent,  of  the 
steam  will  be  condensed. 

: 


62  HEAT  ENGINES 

-7-  8.  Five  pounds  of  steam  at  atmospheric  pressure,  10  Ibs.  of  water  at  60°, 
and  2  Ibs.  of  ice  at  20°  are  mixed  at  atmospheric  pressure.  What  will  be 
the  resulting  temperature? 

9.  Ten  pounds  of  ice  at  10°,  20  Ibs.  of  water  at  60°  and  5  Ibs.  of  steam  at 
atmospheric  pressure  are  mixed  at  atmospheric  pressure.  Find  the  resulting 
temperature  and  condition  of  the  mixture. 

Y-  10.  Twenty  pounds  of  steam  at  atmospheric  pressure,  10  Ibs.  of  water 
at  60°  and  50  Ibs.  of  air  at  100°  are  mixed  together  at  the  pressure  of  the 
steam,  (a)  What  will  be  the  resulting  temperature?  (6)  If  the  steam  is 
not  all  condensed,  determine  what  per  cent,  of  the  steam  will  be  condensed. 

11.  A  mixture  is  made  of  10  Ibs.  of  steam  at  atmospheric  pressure,  5  Ibs. 
of  ice  at  20°,  10  Ibs.  of  water  at  50°,  30  Ibs.  of  air  at  60°.     (a)  What  will  be 
the  temperature  of  the  resulting  mixture?     (6)  What  will  be  the  percentage 
by  weight  of  air,  steam,  and  water  in  the  mixture? 

12.  What  would  be  the  resulting  temperature  and  condition  of  a  mixture 
of  10  Ibs.  of  water  at  40°,  20  Ibs.  of  water  at  60°,  and  8  Ibs.  of  steam  at  5  Ibs. 
pressure?     Mixture  takes  place  at  5  Ibs.  pressure. 

13.  Ten  pounds  of  steam  at  5  Ibs.  pressure,  1  Ib.  of  ice  at  32°,  and  20  Ibs. 
of  water  at  60°  are  mixed  at  5  Ibs.  pressure.     What  will  be  the  temperature 
and  condition  of  the  resulting  mixture? 

,4  14.  Five  pounds  of  ice  at  5°,  10  Ibs.  of  water  at  50°,20  Ibs.  of  air  at80°, 
and  5  Ibs.  of  steam  at  20  Ibs.  pressure  are  mixed  at  the  pressure  of  the  steam. 
Find  the  resulting  temperature  and  condition  of  the  mixture. 

15.  Required  the  temperature  and  condition  of  the  mixture  after  mixing 
10  Ibs.  of  steam  at  a  pressure  of  30  Ibs.  absolute  and   a  temperature  of 
250.3°  F.,  2  Ibs.  of  ice  at  10°  F.,  and  20  Ibs.  of  water  at  40°  F.     Mixture  takes 
place  at  the  pressure  of  the  steam. 

16.  Fifty  pounds  of  air  at  100°,  10  Ibs.  of  steam  at  atmospheric  pressure, 
and  10  Ibs.  of  water  at  60°  are  mixed  at  atmospheric  pressure.     What  is 
the  temperature  of  the  mixture  and  how  much  steam  is  condensed? 

17.  Required  the  final  temperature  and  condition  after  mixing  at  the 
pressure  of  the  air  100  Ibs.  of  air  at  a  temperature  of  500°  and  a  pressure  of 
100  Ibs.  absolute,  and  2  Ibs.  of  steam  at  100  Ibs.  absolute  having  a  quality 
of  98  per  cent. 

)~^is.  Five  pounds  of  steam  at  5  Ibs.  gage  pressure  are  mixed  at  atmospheric 
pressure  with  10  Ibs.  of  water  at  60°.  What  is  the  temperature  and  condition 
of  the  resulting  mixture? 

19.  Thirty  pounds  of  water  at  60°,  10  Ibs.  of  steam  at  115  Ibs.  absolute 
and  a  temperature  of  400°  F.,  and  10  Ibs.  of  ice  at  20°  are  mixed  at  atmos- 
pheric pressure.     What  will  the  resulting  temperature  be?     What  is  the 
condition  of  the  mixture? 

20.  Ten  pounds  of  ice  at  20°  F.,  18  Ibs.  of  water  at  80°,  and  10  Ibs.  steam 
at  75  Ibs.  pressure  and  90  per  cent,  quality,  are  mixed  at  atmospheric  pres- 
sure.    What  is  the  resulting  temperature  and  condition  of  the  mixture? 

21.  Two  pounds  of  steam  at  150  Ibs.  absolute  and  a  temperature  of  400°, 
5  Ibs.  of  ice  at  22°,  and  10  Ibs.  of  water  at  60°  are  mixed  at  atmospheric  pres- 
sure.    Find  the  final  temperature  and  condition  of  mixture. 

22.  Required  the  final  temperature  and  condition  after  mixing  at  atmos- 


CALORIMETERS  AND  MECHANICAL  MIXTURES      63 

pheric  pressure  3  Ibs.  of  ice  at  22°  and  3  Ibs.  of  steam  at  100  Ibs.  pressure  and 
containing  2  per  cent,  moisture. 

23.  Find  the  resulting  temperature  and  condition  of  a  mixture  of  10  Ibs. 
of  steam  at  150  Ibs.  absolute  and  a  temperature  of  400°  F.,  10  Ibs.  of  water 
at  60°  F.,  and  50  Ibs.  of  air  at  112°  F.     Mixture  takes  place  at  atmospheric 
pressure. 

24.  Five  pounds  of  ice  at  0°,  20  Ibs.  of  water  at  75°,  and  15  Ibs.  of  steam 
at  50  Ibs.  absolute  and  95  per  cent,  quality  are  mixed  at  20  Ibs.  absolute. 
What  is  the  resulting  temperature  and  condition  of  the  mixture? 

25.  How  many  pounds  of  water  will  10  Ibs.  of  dry  steam  heat  from  50° 
to  150°  if  the  steam  pressure  is  100  Ibs.  gage? 

26.  If  10  Ibs.  of  steam  at  100  Ibs.  gage  raises  93  Ibs.  of  water  from  50° 
to  140°,  what  per  cent,  of  moisture  is  in  the  steam,  radiation  being  zero? 

27.  A  pound  of  steam  and  water  occupies  3  cu.  ft.  at  110  Ibs.  absolute 
pressure.     What  is  the  quality  of  the  steam? 


CHAPTER  V 
COMBUSTION  AND  FUELS 

41.  Coal  Analysis. — The  source  of  heat  which  is  used  to  pro- 
duce steam  in  a  boiler  is  the  fuel.  The  principal  ingredients  of 
all  fuels  are  carbon  and  hydrogen. 

For  the  purpose  of  making  comparison  between  the  product  of 
various  mines,  and  to  determine  the  relative  value  of  these  fuels 
for  different  uses,  coal  is  subjected  to  an  ultimate  and  a  proximate 
analysis,  and  it  is  tested  in  a  coal  calorimeter  to  ascertain  its 
calorific  or  heating  value. 

In  the  ultimate  analysis  the  proportions  of  carbon,  hydrogen, 
oxygen,  nitrogen  and  sulphur  are  determined. 

In  the  proximate  analysis  determinations  are  made  of  the 
amounts  of  moisture,  volatile  matter,  fixed  carbon  and  ash.  The 
volatile  gases  are  hydrocarbons  such  as  marsh  or  olefiant  gas, 
pitch,  tar  and  naphtha.  All  of  these  must  be  distilled  from  the 
fuel  before  being  burned.  The  fixed  carbon  and  ash  are  left 
after  all  the  volatile  gases  have  been  driven  off.  The  ash  con- 
sists of  the  incombustible  material  which  remains  after  the  fuel 
has  been  completely  burned. 

It  should  be  noted  that  the  term  " proximate"  does  not  mean 
that  the  analysis  is  only  "  approximate,"  the  facts  being  actually 
to  the  contrary. 

The  analyses  are  made  of  air-dried  coal.  Therefore,  the 
various  percentages  of  carbon,  hydrogen,  etc.,  in  the  ultimate 
analysis,  and  of  fixed  carbon,  ash,  etc.,  in  the  proximate  analysis 
of  "coal  as  received,"  "moisture  free"  or  "dry  coal,"  and  "coal, 
moisture  and  ash  free,"  as  given,  for  example,  in  the  U.  S.  Bureau 
of  Mines  Bulletin  No.  22  on  the  "Analyses  of  Coals  in  the 
United  States,"  "were  not  obtained  directly  but  were  calculated 
from  the  values  obtained  by  the  analyses  of  air-dried  coal." 

64 


COMBUSTION  AND  FUELS  65 

"Calculations  from  'Air  Dried'  to  'Moisture  Free'  Condition  "* 
'Air  dried'  condition  'Moisture  free'  condition 

100 
Volatile  matter  <  100  _  moisture  =  volatile  matter 

100 
Fixed  carbon  X  100  _  molstur^  =  fixed  carbon 

Ash  X  100  -Toisture  =  ash 

100 
SulPhur  <  100  -  moisture  =  sulphur 

100 
Hydrogen  (-1/9  moisture)  X  100  _  moisture  =  hydr°gen 

100 

Carbon  <  100  -  moisture  =  carbon 

Nitrogen  X  100  -Toisture  =  nitrogen 

100 
Oxygen  (-  8/9)  moisture)       <         ^  =  oxy^en 


100 
Calorific  value  <  100  _  moisture  =  calorific  value 

"The  analyses  are  calculated  to  the  'moisture  and  ash  free'  basis  by 
taking  100—  (moisture  +  ash)  as  a  divisor  and  proceeding  otherwise 
exactly  as  in  the  calculation  to  the  'dry  coal'  or  'moisture  free'  basis. 

"The  air-drying  loss  of  a  mine  sample  indicates  to  some  degree  the 
loss  in  weight  after  mining  from  the  evaporation  of  loosely  retained 
moisture.  The  analysis  of  the  coal  'as  received'  shows  the  actual 
composition  of  the  coal  in  the  mine.  After  the  coal  has  left  the 
mine  its  moisture  content  lies  between  the  limits  of  coal  'as  received.  » 
and  coal  'air  dried.' 

"The  analysis  on  a  'moisture  free'  basis  represents  the  composition 
of  the  coal  after  drying  at  221°  F.  (105°  C.). 

"The  analysis  stated  on  a  'moisture  and  ash  free'  basis  represents 
approximately  the  heating  value  and  composition  of  the  dry  organic 
matter.  This  relation  seems  to  be  fairly  constant  for  the  same  coal 
bed  in  certain  districts,  especially  in  the  Appalachian  region.  Com- 
parison of  numerous  analyses  shows  that  the  'moisture  and  ash  free' 
calorific  values  of  different  samples  from  the  same  mine  and  bed 
usually  agree  closely,  provided  the  proportion  and  the  character  of 
the  ash  and  the  sulphur  do  not  vary  greatly. 

"For  the  commercial  valuation  of  coals  a  proximate  analysis  and  a 
calorific  value  determination  are  usually  sufficient.  Moisture  and  ash 
are  of  importance;  they  not  only  displace  their  own  weights  of  com- 
bustible matter,  but  the  evaporation  of  the  moisture  wastes  heat. 

*  U.  S.  Bureau  of  Mines  Bulletin  No.  22. 
5 


66  HEAT  ENGINES 

A  high  percentage  of  ash  increases  the  cost  of  handling  coal  in  a 
power  plant  and  decreases  the  efficiency  of  the  furnace. 

"The  ratio  of  the  volatile  matter  to  the  fixed  carbon  indicates  in  a  way 
the  type  of  furnace  best  adapted  for  burning  a  coal  with  maximum 
efficiency. 

"The  smokeless  combustion  of  coal  containing  a  low  percentage  of 
volatile  matter  is  not  difficult  in  furnaces  of  ordinary  types,  but  to  burn 
a  high  volatile  coal  without  smoke  requires  a  suitably  designed  furnace. 
A  high  percentage  of  sulphur  is  undesirable  in  coal  used  for  the  manufac- 
ture of  coke  and  gas.  For  ordinary  steaming  purposes  sulphur  is  not  a 
serious  drawback  unless  associated  with  elements,  such  as  iron  or  lime, 
that  promote  clink ering." 

42.  Heat  of  Combustion. — The  term  combustion  as  applied 
here  refers  to  the  union  of  oxygen  with  some  other  substance  pro- 
ducing heat.  The  perfect  combustion  of  ordinary  fuel  should 
result  in  carbon  dioxide,  nitrogen,  water  vapor,  and  a  trace  of 
sulphur  dioxide. 

"The  calorific  power,  or  heating  value,  of  a  fuel  is  the  total 
amount  of  heat  developed  by  the  complete  combustion  of  a  unit 
weight  of  fuel."  The  calorific  power  as  determined  by  a  calo- 
rimeter is  the  higher  heating  value.  When  a  fuel  is  burned  water 
vapor  is  formed,  and  this  will  be  condensed  only  when  the 
temperature  falls  below  the  boiling  point.  So  long  as  this 
water  remains  in  the  form  of  vapor,  the  heat  necessary  to  main- 
tain it  as  such,  i.e.,  the  latent  heat  of  steam  at  atmospheric 
pressure  times  the  weight  of  vapor,  is  unavailable  for  use. 

The  difference  between  the  higher  heating  value  and  this 
latent  heat  is  called  the  lower  heating  value.  This  is  the  "  available 
calorific  value"  in  nearly  all  cases.  For  example,  in  a  boiler  plant 
the  temperature  of  the  stack  gases,  and  in  a  gas  engine  the 
temperature  of  the  exhaust,  are  both  above  the  temperature  of 
the  boiling  point  of  water,  and  therefore  the  heat  actually  avail- 
able for  use  in  either  case  is  the  lower  heating  value  of  the  fuel. 

The  heat  given  off  per  pound  by  the  elements  ordinarily  met 
with  in  fuels,  together  with  the  air  required  for  combustion  and 
the  combining  volumes  and  weights,  are  shown  in  Table  IX. 


COMBUSTION  AND  FUELS 


67 


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68  HEAT  ENGINES 

When  a  coal  is  analyzed  the  percentage  of  hydrogen  shown 
includes  not  only  the  free  hydrogen  in  the  sample  but  also  that 
which  existed  in  combination  with  the  oxygen  in  the  form  of 
water  (for  all  the  oxygen  in  the  coal  will  be  united  with  hy- 
drogen). As  16  parts  by  weight  of  oxygen  unite  with  2  parts 
of  hydrogen,  the  weight  of  hydrogen  which  was  in  combination 
with  the  oxygen  will  be  equal  to  one-eighth  the  total  weight  of 
oxygen.  The  balance  of  the  hydrogen  is  available  for  producing 
heat,  and  in  determining  the  heat  value  of  a  fuel,  the  number 
of  B.T.U.  in  the  coal  may  be  found  from  the  ultimate  analysis 
by  the  following  formula: 
Heat  value  of  fuel  in  B.T.U.  per  pound 


=  14,650  C  +  62,100  (H  --  g    +  4000  S,  (1) 

where  the  symbols  C,  H,  0  and  S  represent  the  weights  of  car- 
bon, hydrogen,  oxygen  and  sulphur  in  1  Ib.  of  the  fuel.  This  is 
called  Du  Long's  formula. 

The  heat  value  obtained  from  equation  (1)  is  only  an  approxi- 
mate result,  and  where  greater  accuracy  is  desired  it  is  necessary 
actually  to  test  the  coal  experimentally  in  a  coal  calorimeter. 

43.  Coal  Calorimeters.  —  One  form  of  calorimeter  very  com- 
monly used  for  determining  the  heating  value  of  solid  fuels 
is  the  Mahler  Bomb  Calorimeter.  This  consists  of  a  strong 
steel  vessel  into  which  a  known  weight  (usually  1  gram)  of 
finely  powdered  air-dried  coal  is  introduced.  This  coal  is  placed 
in  a  platinum  cup  or  dish  suspended  from  the  cover  of  the  bomb 
by  a  wire  electrode.  Another  wire  passes  through  the  cover, 
although  well  insulated  from  it,  and  extends  down  into  the  coal. 
The  cover  is  then  screwed  down  tight  and  the  bomb  charged  with 
oxygen  to  a  pressure  of  from  150  to  250  Ibs.  This  allows  a 
considerable  excess  of  oxygen  over  that  theoretically  required 
for  the  combustion  of  the  coal.  After  the  bomb  has  been 
charged  it  is  placed  in  a  vessel  containing  a  known  weight  of  water 
and  an  electric  current  is  passed  through  the  wire  electrodes, 
igniting  the  coal.  While  the  combustion  is  going  on,  the  water 
in  the  containing  vessel  is  kept  thoroughly  stirred  by  the  ap- 
paratus. 

The  rise  in  temperature  of  the  water  is  carefully  noted,  and 
after  making  allowances  for  radiation,  the  heat  generated  by  the 
electric  current,  etc.,  the  heating  value  of  the  coal  can  be  com- 


COMBUSTION  AND  FUELS 


69 


puted,  since  the  heat  gained  by  the  water  must  equal  the  heat 
given  up  by  the  coal  (after  the  allowances  just  mentioned  have 
been  made). 

Another  type  which  has  found  considerable  use  in  cases  where 
it  is  not  convenient  to  secure  a  supply  of  oxygen  under  pressure, 


N 


FIG.  10. — Parr  coal  calorimeter. 


FIG.  11. — Cartridge  in 
Parr  calorimeter. 


is  the  Parr  Calorimeter  shown  in  Fig.  10.  This  is  simpler  to 
use  than  a  bomb  calorimeter,  but  the  results  obtained  are  not  as 
accurate.  The  charge  consists  of  .004  of  a  pound  of  finely 
powdered  coal  and  eighteen  times  as  much  by  weight  of  sodium 
peroxide  to  supply  the  oxygen  for  combustion.  After  the  charge 
has  been  placed  in  the  cartridge,  Fig.  11,  and  A,  Fig.  10,  and  the 
cover  has  been  tightly  screwed  down,  it  should  be  thoroughly 


70  HEAT  ENGINES 

mixed  by  shaking.  The  calorimeter  is  then  immersed  in  a  vessel 
containing  water  and  a  short  piece  of  white  hot  wire  is  dropped 
in  the  top  of  the  long  neck  and  a  blow  on  the  upper  end  opens 
the  valve  at  M,  Fig.  10,  and  allows  the  wire  to  drop  into  the 
charge  igniting  it.  The  water  is  stirred  by  fins  attached  to  the 
sides  of  the  cartridge  which  is  turned  on  a  pivot  bearing  at  the . 
bottom  by  a  belt  run  by  an  electric  motor.  The  rise  in  tempera- 
ture of  the  water  due  to  the  combustion  of  the  coal  is  carefully 
noted.  After  making  allowances  for  the  heat  radiated,  the  heat 
given  up  by  the  combustion  of  the  sodium  peroxide  and  by  the 
wire  used  for  ignition,  the  heating  value  of  the  coal  is  found  just 
as  in  the  case  of  the  bomb  calorimeter. 

44.  Air  Required  for  Combustion. — The  oxygen  furnished  to 
the  fuel  in  order  to  burn  it  is  obtained  from  the  air.  Air  is  a 
mechanical  mixture  containing  by  weight  23  per  cent,  oxygen 
and  77  per  cent,  nitrogen,  and  by  volume  21  per  cent,  oxygen  and 
79  per  cent,  nitrogen.  The  oxygen  only  is  used  in  the  combustion 
of  the  fuel,  the  nitrogen  being  an  inert  gas  and  having  no 
chemical  effect  upon  the  combustion. 

For  the  complete  combustion  of  1  Ib.  of  hydrogen  there 
is  required  8  Ibs.  of  oxygen,  and  for  the  complete  combustion 
of  1  Ib.  of  carbon  to  carbon  dioxide  there  is  required  32  -f-  12  = 
2.66  Ibs.  of  oxygen.  For  each  pound  of  hydrogen  there  will 

Q 

be  required  —^  =  34.8  Ibs.  of  air,  and  for  each  pound  of  carbon 

.  Zo 

f\     C*C* 

00  •=  11.6  Ibs.  of  air  to  produce  combustion. 

.  Zo 

As  has  already  been  stated,  the  oxygen  in  the  fuel  unites  with 
its  equivalent  of  hydrogen  to  form  water  and  in  determining 
the  weight  of  air  theoretically  required  for  combustion  this  hy- 
drogen should  be  disregarded.  The  air  required  for  the  com- 
plete combustion  of  any  fuel  may  then  be  found  from  its  analysis 
by  the  following  expression: 

Weight  of  air  per  pound  of  fuel 
=  11.6  C  +  34.8  (H  --  |)  +  4.35  S  (2) 

In  equations  (1)  and  (2)  it  has  been  assumed  that  each  atom 
of  hydrogen  and  carbon  comes  in  contact  with  a  proper  pro- 
portion of  oxygen.  In  actual  practice  this  condition  does  not 
exist  and  an  excess  of  air  is  furnished  in  order  to  insure  com- 


COMBUSTION  AND  FUELS  71 

plete  combustion.  Theoretically  most  coals  require  for  com- 
plete combustion  approximately  12  Ibs.  of  air.  In  actually 
burning  coal  under  a  boiler  with  natural  draft  we  find  that 
the  coal  requires  about  24  Ibs.  of  air  per  pound  of  coal.  For 
forced  draft  there  is  usually  required  about  18  Ibs.  per  pound 
of  coal.  If  insufficient  air  is  admitted  to  the  fire,  only  a  por- 
tion of  the  carbon  will  unite  with  the  oxygen  to  form  C02,  the 
balance  forming  CO. 

In  the  actual  operation  of  a  boiler  plant,  one  of  the  most 
important  considerations  is  the  admission  of  a  proper  quantity 
of  air  to  the  fire.  As  will  be  seen  later,  the  less  the  quantity 
of  air  given  to  the  fire  the  better  the  efficiency  of  combustion, 
provided  enough  air  enters  so  that  all  the  carbon  is  burned  to 
C02. 

45.  Smoke. — Smoke  is  unburned  carbon  in  a  finely  divided 
state.     The  amount  of  carbon  carried  away  by  the  smoke  is 
usually  small,  not  exceeding   1  per  cent,  of  the  total  carbon 
in  the  coal.     Its  presence,  however,  often  indicates  improper 
handling  of  the  boiler,  which  may  result  in  a  much  larger  waste 
of  fuel.     Smoke  is  produced  in  a  boiler  when  the  incandescent 
particles  of  carbon  are  cooled  before  coming  into  contact  with 
sufficient  oxygen  to  unite  with  them.     It  is  necessary  that  the 
carbon  be  in  an  incandescent  condition  before  it  will  unite 
with  the  oxygen.     Any  condition  of  the  furnace  which  results 
in  carbon  being  cooled  below  the  point  of  incandescence  before 
sufficient  oxygen  has  been  furnished  to  unite  with  it,  will  result 
in  smoke.     Smoke  once  formed  is  very  difficult  to  ignite,  and 
the  boiler  furnace  must  be  handled  so  as  not  to  produce  smoke. 
Fuels  very  rich  in  hydrocarbons  are  most  apt  to  produce  smoke. 
When  the  carbon  gas  liberated  from  the  coal,  is  kept  above  the 
temperature  of  ignition  and  sufficient  oxygen  for  its  combus- 
tion added,  it  burns  with  a  red,  yellow,  or  white  flame.     The 
slower  the  combustion  the  larger  the  flame.     When  the  flame 
is  chilled  by  the  cold  heating  surfaces  near  it  taking  away  heat 
by  radiation,  combustion  may  be  incomplete,  and  part  of  the 
gas  and  smoke  pass  off  unburned.     If  the  boiler  is  raised  high 
enough  above  the  grate  so  as  to  give  room  for  the  volatile  matter 
to  burn  and  not  strike  the  tubes  at  once,  the  amount  of  smoke 
given  off  and  of  coal  used  will  both  be  reduced. 

46.  Analysis  of  Flue  Gases. — In  all  large  power  houses  and 
carefully  conducted  power  plants  the  flue  gases  leaving  the 


72 


HEAT  ENGINES 


boilers  are  analyzed  from  time  to  time.  In  some  cases  records 
are  kept,  by  an  automatic  device,  of  the  percentage  of  carbon 
dioxide  in  the  flue  gases.  In  analyzing  the  flue  gases  it  is  cus- 
tomary to  use  some  modification  of  the  Orsat  apparatus.  This 
consists  of  three  pipettes,  a  measuring  tube,  and  a  wash  bottle, 
as  shown  in  Fig.  12.  The  first  pipette  D  contains  a  saturated 
solution  of  potassium  hydrate  and  absorbs  CO2,  the  second 
pipette  E  contains  potassium  pyrogallate  and  absorbs  O,  and 
the  third  pipette  F  contains  cuprous  chloride  and  absorbs  CO. 
The  gas  is  passed  through  the  pipettes  in  the  order  named,  and 


FIG.  12. — Orsat  apparatus. 

the  remainder  is  assumed  to  be  nitrogen.  The  readings  obtained 
from  this  apparatus  give  the  per  cent,  composition  of  the  gases  by 
volume. 

The  following  directions  will  show  how  the  reagents  used  in 
the  Orsat  apparatus  are  prepared. 

Potassium  Hydrate. — (1)  For  the  determination  of  C02, 
dissolve  500  grams  of  the  commercial  hydrate  in  1  liter  of  water. 

1  c.c.  of  this  solution  will  absorb  40  c.c.  of  C02. 

(2)  For  the  preparation  of  potassium  pyrogallate  for  use 
in  case  the  per  cont.  of  oxygen  is  high,  dissolve  120  grams  of  the 
commercial  hydrate  in  100  c.c  of  water. 


COMBUSTION  AND  FUELS  73 

Potassium  Pyrogallate. — Put  5  grams  of  the  solid  pyrogailic 
acid  in  a  funnel  placed  in  the  neck  of  the  pipette  E,  and  pour 
over  this  100  c.c.  of  potassium  hydrate,  solution  (1)  or  (2).  So- 
lution (1)  may  be  used  in  case  there  is  not  more  than  25  per 
cent,  of  0  in  the  gas.  Otherwise  solution  (2)  must  be  used  or  CO 
may  be  given  up. 

1  c.c  of  this  solution  absorbs  2  c.c.  of  0. 

Cuprous  Chloride. — Pour  from  J  to  \  an  inch  of  copper  scale 
into  a  2-liter  bottle  and  also  place  in  the  bottle  a  number 
of  long  pieces  of  copper  wire.  Then  fill  the  bottle  with  hydro- 
chloric acid  of  1.10  sp.  gr.  (1  part  muriatic  acid  to  1  part  water). 
Let  the  bottle  stand,  shaking  it  occasionally  until  the  solution 
becomes  colorless.  Then  pour  the  liquid  into  the  pipette  F, 
which  is  filled  with  copper  wires. 

1  c.c  of  this  solution  will  absorb  from  1  to  2  c.c  of  CO. 

Example. — A  stack  gas  shows  the  following  analysis:  C02,  12  per 
cent. ;  CO,  1  per  cent. ;  0,  7  per  cent. ;  N,  80  per  cent.  Find  the  air  used 
in  burning  a  pound  of  coal,  if  the  coal  contains  C,  80  per  cent.;  H,  4  per 
cent.;  0,  2  per  cent. 

Solution. — 

Vol.  in  100  Density      Weight 
cu.  ft. 

Carbonic  acid,  C02     12    X       .12341    =    1.481 

Carbonic  oxide,  CO    1    X       .07806    =       .078 

Oxygen,  0    7    X       .08928    =       .625 

One  pound  of  carbon  dioxide  contains  i\  of  a  pound  of  oxygen,  and 
1  Ib.  of  carbonic  oxide  contains  y  of  a  pound.  The  weight  of  the  oxgyen 
in  100  cu.  ft.  of  the  flue  gases  would  therefore  be: 

In  carbonic  acid    A  X  1.481  =  1.077 

In  carbonic  oxide    *  X    .078  =  .045 

Free  oxygen     =  .625 

Total  weight  of  oxygen  =  1.747*  pounds 
and  the  weight  of  the  carbon  would  be: 

In  carbonic  acid     A  X  1.481  =    .404 

In  carbonic  oxide   2i  X     .078  =    X)33 

Total  weight  of  carbon    437  pounds 

Air  contains  23  per  cent,  of  oxygen  by  weight;  hence  the  pounds  of  air 
required  to  burn  .437  Ibs.  of  carbon  would  be 

1.747  ^  .23  =  7.6, 


74  HEAT  ENGINES 

and  the  pounds  of  air  to  burn  1  Ib.  of  carbon  under  the  conditions  of 
the  flue  gases  would  be 

7.6  -f-  .437  =  17.4 

The  pounds  of  air  used  to  burn  a  pound  of  coal  of  the  given  analysis 
would  be 


17.4   C  +  34 

=  17.4  X  .80  +  34.8  (0.4  -  ~\  =  13.92  +  1.31 
=  15.23  Ibs. 

It  should  be  noted  here  that  in  this  solution  the  weight  of  air  theoretic- 
ally required  to  burn  the  hydrogen  has  been  added  to  the  weight  actually 
required  to  burn  the  carbon  as  shown  by  the  stack  gas  analysis.  While 
this  is,  of  course,  not  exactly  correct,  it  is  approximately  so,  and  the  error 
is  slight,  as  the  amount  of  air  used  to  burn  the  hydrogen  is  small  as  com- 
pared with  the  total  amount  required. 

The  above  results  are  such  as  might  be  expected  in  a  boiler  plant  using 
induced  draft. 

47.  Theoretical  Temperature  of  Combustion.  —  If  the  total 
and  specific  heats  of  the  materials  of  a  given  coal  are  known,  the 
temperature  that  might  result  from  their  combustion  may  be 
approximately  calculated. 

The  calculated  temperatures  are  often  very  much  higher 
than  can  be  obtained  in  practice,  this  being  probably  due  to 
the  fact  that  the  specific  heat  of  the  products  of  combustion 
is  very  much  larger  at  the  high  temperatures,  and  also  to  the 
fact  that  carbon  and  oxygen  will  no  longer  unite  above  a  given 
temperature,  probably  about  3500°  Fahrenheit.  . 

Example.  —  Assume  the  following  composition  of  coal:  Carbon,  75 
per  cent.;  hydrogen,  5  per  cent.;  oxygen,  3  per  cent.;  nitrogen,  2  per, 
cent.;  the  ash  and  sulphur  may  be  disregarded.  Find  the  theoretical 
and  actual  rise  in  temperature  of  the  products  of  combustion. 

Solution.  —  A  coal  of  the  above  composition  has  a  heat  value  of  13,860 
B.T.U.  The  theoretical  amount  of  air  required  to  burn  1  Ib.  of  it  is 
10.62  Ibs.  10.62  Ibs.  of  air  contain  10.62  X  .77  =  8.18  Ibs.  nitrogen, 
to  which  must  be  added  the  .02  Ibs.  of  nitrogen  in  the  coal,  giving  us 
a  total  of  8.2  Ibs.  nitrogen. 

Total  C02  formed  =  .75  X  3.66  =  2.745  Ibs.  }  (See 

Total  H20  formed  =  .05  X  9        =    .45    Ibs.  J          Table  IX) 

The  thermal  units  required  to  raise  the  products  of  combustion 
through  1°  would  be 


COMBUSTION  AND  FUELS  75 

Sp.  ht.    B.T.U. 

Carbonic  acid 2.75  X  .217  =    .596 

Water  vapor 45  X  .460  =    .207 

Nitrogen 8.2    X  .244  =  2.000 


Total 2.803 

The  theoretical  rise  in  temperature  of  the  products  of  combustion 
would  be 

13,860  -h  2.8  =  4950° 

In  the  actual  operation  of  a  boiler  it  is  found  necessary  to  add  50 
to  100  per  cent,  more  air  than  is  required  for  combustion.  This  addi- 
tional air,  as  the  following  calculation  shows,  materially  reduces  the 
theoretical  temperature  of  combustion.  Assuming  100  per  cent. more 
to  be  required,  there  would  then  be  added  10.62  additional  pounds  of 
air.  The  heat  to  raise  this  1  degree  would  be 

10.62  X  .2375  =  2.522 
Add  for  undiluted  products 2.803 

Total  B.T.U.  per  degree 5.325 

The  theoretical  rise  in  temperature  would  be,  then, 
13,860  -^  5.325  =  2600° 

This  is  more  nearly  the  temperature  obtained  in  a  boiler  plant  with 
hand  firing. 

If  the  temperature  of  the  boiler  room  is  given,  the  final 
temperature  of  the  products  of  combustion  may  be  found  by 
adding  to  this  temperature  the  rise  in  temperature  as  found 
above,  the  assumption  being  made  that  the  temperature  of 
the  coal  is  the  same  as  that  of  the  boiler  room. 

In  boilers  operated  by  automatic  stokers,  temperatures  in  the 
fire  of  over  3000°  F.  have  been  observed.  Such  temperatures  are 
usually  obtained  when  the  boilers  are  being  crowded  to  their 
full  capacity  and  their  operation  is  being  given  careful  atten- 
tion, especially  with  reference  to  the  amount  of  air  admitted  to 
the  furnace. 

48.  Fuels. — Fuels  may  be  divided  into  three  general  classes, 
solid,  liquid,  and  gaseous. 

The  larger  proportion  of  the  fuels  used  are  in  solid  form. 
The  principal  solid  fuels  are  wood,  peat,  lignite,  and  coal.  Coal 
may  be  divided  into  three  principal  kinds,  anthracite,  semi- 
bituminous  and  bituminous  coal. 


76 


HEAT  ENGINES 


The  liquid  fuels  are  usually  some  of  the  mineral  oils,  generally 
unrefined  petroleum.  In  some  gas  plants  liquid  tar  is  used. 

The  most  commonly  used  gaseous  fuel  is  natural  gas,  but 
there  are  a  good  many  plants  using  gas  which  is  a  waste  product 
from  a  manufacturing  operation.  In  the  steel  mills  the  "down 
comer"  gases  from  the  blast  furnaces  are  often  used  as  a  fuel 
for  the  steam  boilers.  Coke-oven  gases  are  similarly  used. 
In  some  cases  the  coal  is  distilled  in  a  gas  producer,  and  this 
producer  gas  used  as  a  fuel. 

49.  Woods. — Woods  may  be  divided  into  two  general  classes, 
soft  and  hard.  The  commonest  hard  woods  are  oak,  hickory, 
maple,  beech,  and  walnut.  The  commonest  soft  woods  are 
pine,  elm,  birch,  poplar,  and  willow.  When  first  cut,  wood 
contains  about  50  per  cent,  of  moisture,  but  after  being  dried 
this  is  reduced  from  10  to  20  per  cent.  The  following  table 
gives  the  chemical  composition  and  heat  value  of  some  of  the 
more  common  woods.  (From  Poole's  Calorific  Value  of  Fuels.) 


TABLE  X. — CALORIFIC  VALUE  OF  WOODS 


Name 

C 

H 

O 

N 

Ash 

B.T.U 
per  Ib. 

combustible 

Ash 

49  2 

6  3 

43.9 

.07 

.57 

8480 

Beech                 

49.0 

6.1 

44.2 

.09 

.57 

8590 

Birch 

48  9 

6  0 

44.7 

10 

29 

8590 

Elm                          .      ..... 

48.9 

6.2 

44.3 

.06 

50 

8510 

Oak  

50.2 

6.0 

43.4 

.09 

.37 

8320 

Pine 

50  3 

6.2 

43.1 

.04 

37 

9150 

In  boiler  tests  a  pound  of  wood  is  usually  assumed  as  equal 
to  .4  of  a  pound  of  coal. 

50.  Peat. — Peat  is  an  intermediate  between  wood  and  coal. 
It  is  formed  from  the  immense  quantity  of  rushes,  sedges,  and 
mosses  that  grow  in  the  swampy  regions  of  the  temperate  zone. 
These  in  the  presence  of  heat  and  moisture  are  subject  to  a 
chemical  change  which  leaves  behind  the  hydrocarbons,  fixed 
carbon,  and  70  to  80  per  cent,  of  moisture.  It  is  usually  cut 
in  blocks  and  air  dried.  Good  air-dried  peat  contains  about 
60  per  cent,  of  carbon,  6  per  cent,  of  hydrogen,  31  per  cent, 
of  oxygen  an,d  nitrogen,  and  3  per  cent,  of  ash.  The  following 
table  gives  the  heat  value  of  some  of  the  different  peats: 


COMBUSTION  AND  FUELS 


77 


TABLE  XI.  CALORIFIC  VALUE  OP  PEATS 


Location 

Fixed 
carbon 

Volatile 
matter 

Ash 

B.T.U.  per  Ib. 
combustible 

Northern  Michigan  

4  4 

11  000 

Southern  Michigan 

33  3 

61   2 

5  5 

8  900 

Southern  Michigan  

29.0 

68  5 

2  3 

9  500 

New  York 

29  2 

65  6 

8  25 

10200 

Wisconsin  

27.6 

60  5 

11  8 

8  250 

51.  Lignite  Coal. — Lignite  is  coal  of  very  recent  formation, 
and  its  analysis  is  similar  to  peat.  It  usually  resembles  wood  in 
appearance,  and  is  of  brownish  color.  It  is  uneven  of  fracture 
and  of  a  dull  luster.  It  is  found  quite  generally  west  of  the 
Mississippi  River.  The  composition  is  given  in  the  following 
table: 


TABLE  XII.   CALORIFIC  VALUE  OF  LIGNITES 


Location 

Fixed 
carbon 

Volatile 
matter 

Ash 

B.T.U.  per  Ib. 
combustible 

California  

9063 

Colorado  

46 

32.7 

2  74 

11,360 

52.  Bituminous  Coal. — Coals  that  contain  over  20  per  cent, 
volatile  matter  are  usually  classed  as  bituminous  coals.  Bitu- 
minous coals  are  divided  into  coking,  non-coking,  and  cannel 
coals. 

" Coking  coal"  is  a  term  used  in  reference  to  coals  that  fuse 
together  on  being  heated  and  become  pasty  These  coals  are 
used  in  gas  manufacture,  and  are  very  rich  in  hydrocarbons. 
Non-coking  coals  are  free  burning  and  the  lumps  do  not  fuse 
together  on  being  heated.  " Jackson  Hill"  is  an  example  of 
this  kind  of  coal.  Cannel  coal  is  very  rich  in  carbon,  ignites 
readily,  and  burns  with  a  bright  flame.  It  is  very  homogeneous, 
breaks  without  any  definite  line  of  fracture,  and  has  a  dull, 
resinous  luster.  It  is  very  valuable  as  a  gas  coal  so  that  it  is 
little  used  for  steaming  purposes. 

The  principal  bituminous  coals  used  are  mined  in  Ohio, 
West  Virginia,  Pennsylvania,  and  Illinois.  The  following  table 
gives  the  properties  of  the  commonest  varieties  of  the  bituminous 
coals  used  for  steaming  purposes: 


78 


HEAT  ENGINES 


TABLE  XIII.  CALORIFIC  VALUE  OF  BITUMINOUS  COALS 


Location 

Fixed 
carbon 

Volatile 
matter 

Ash 

B.T.U.  per  Ib. 
combustible 

Water 

Illinois: 
Big  Muddy              .    . 

53  7 

30  1 

9  2 

13,610 

Streator  

44.0 

39.2 

12.3 

13,690 

4.5 

Wilmington 

44  9 

36  8 

13  3 

14,050 

13  3 

Michigan: 
Saginaw 

6  1 

13,470 

Ohio: 
Brier  Hill... 

59.1 

36.4 

4  5 

14,200 

Hocking  Valley  

49.1 

36.1 

8.5 

13,980 

6.4 

Jackson  ....              ... 

54  6 

34.3 

7  0 

13,955 

4  1 

Pennsylvania: 
Pittsburg   No.   8 

54  6 

35.5 

9  9 

14,200 

Turtle  Creek  

56.6 

34.4 

8.0 

15,080 

1.0 

Youghiogheny               .  . 

54  7 

32  6 

12  7 

15,000 

West  Virginia-' 
Clover  Hill 

56  8 

31  7 

10  1 

14,265 

Thacker  

56.2 

35.5 

6.8 

15,240 

53.  Semi-Bituminous. — This  is  a  softer  coal  than  anthracite, 
but  in  appearance  it  looks  like  the  latter.  It  is  lighter  than 
anthracite  and  burns  more  rapidly,  and  is  a  valuable  coal  where 
it  is  necessary  to  keep  a  very  intense  heat.  Its  composition  is 
given  in  the  following  table : 

TABLE  XIV.  CALORIFIC  VALUE  OF  SEMI-BITUMINOUS  COALS 


Location 

Fixed  , 
carbon 

Volatile 
matter 

Ash 

B.T.U.  per  Ib. 
combustible 

Blassburg,  Pa               

73  0 

15  0 

11.0 

13,500 

Cumberland,  Md  

80.8 

13.0 

5.0 

16,320 

Pocahontas,  W.  Va             .  .  . 

74  5 

18  1 

6  6 

15,740 

A  semi-bituminous  coal  should  not  contain,  usually,  more 
than  20  per  cent,  volatile  matter  as  compared  with  the  fixed 
carbon. 

54.  Anthracite. — This  coal  ignites  very  slowly  and  burns  at 
a  high  temperature.  Its  principal  component  is  fixed  carbon. 
Consequently  it  gives  off  almost  no  smoke  and  the  flame  is  very 
short.  Owing  to  its  smokeless  burning,  it  is  almost  all  consumed 
for  domestic  purposes.  Nearly  all  anthracite  used  in  this  country 
comes  from  Pennsylvania.  An  anthracite  coal  should  contain  not 


COMBUSTION  AND  FUELS  79 

less  than  92  per  cent,  of  fixed  carbon  as  compared  with  the  volatile 
matter.  The  following  is  a  table  of  the  composition  of  various 
anthracite  coals: 

TABLE  XV.  CALORIFIC  VALUE  OF  ANTHRACITE  COALS 


Location 

Fixed 
carbon 

Volatile 
matter 

Ash 

B.T.U.  per  Ib. 
combustible 

Lackawam 
Lykens  Va 
Scran  ton 

la 

84.0 
81.0 

84.4 

5.0 
5.0 
6.5 

11.0 
14.0 
9.0 

13,900 
13,650 
13,800 

Hey     

55.  Efficiency  of  Fuels. — The  commercial  value  of  a  fuel  is 
determined  by  the  number  of  pounds  of  water  it  will  evaporate 
into  steam  per  hour  from  and  at  212°.  This,  however,  involves 
the  efficiency  of  the  boiler,  so  that  to  compare  fuels  in  actual 
use,  they  should  be  burned  in  the  same  boiler.  In  practice 
the  value  of  a  fuel  in  any  given  plant  is  affected  by  the  form 
and  character  of  the  furnace,  the  amount  of  air  supplied,  and  the 
intensity  of  the  draft.  There  are,  in  fact,  so  many  variables 
entering  into  the  problem  that  it  is  difficult  to  make  an  accurate 
comparison  of  the  value  of  the  different  coals. 

It  is  easy  to  burn  either  anthracite  or  semi-bituminous  coal 
in  almost  any  boiler.  For  bituminous  coals  containing  less 
than  40  per  cent,  volatile  matter,  plain  grate  bars  with  a  fire- 
brick arch  over  the  fire  give  very  good  results.  With  coals 
containing  over  40  per  cent,  volatile  matter,  it  is  desirable  to 
use  some  form  of  furnace  arranged  so  that  the  gases  are  mixed 
with  warm  air,  and  with  these  a  large  combustion  chamber 
should  be  provided. 

The  commercial  results  obtained  from  a  given  coal  are  usually 
determined  by  the  cost  to  evaporate  1000  Ibs.  of  water  into  steam 
from  and  at  212°.  This  cost  varies  from  10  cents  to  18  cents. 
Where  the  principal  cost  of  the  coal  is  in  the  freight  rate,  it  is 
usually  more  economical  to  burn  a  good  grade  of  coal  than  a 
cheap  grade. 

PROBLEMS 

1.  An  anthracite  has  the  following  composition:  C,  90  per  cent.;  H, 
2  per  cent.;  O,  2  per  cent.     Find  the  heating  value  of  the  coal. 

2.  A  semi-bituminous  coal  has  the  following  composition:  C,   80  per 
cent.;  H,  5  per  cent.;  O,  3  per  cent.     Find  the  heat  units  in  the  coal. 


80  HEAT  ENGINES 

v 

3.  A  Pennsylvania  bituminous  coal  contains:  C,  75  per  cent.;  H,  5  per 
cent.;  O,  12  per  cent.     Find  the  heat  value  of  the  coal  and  the  air  required  to 
burn  1  Ib. 

4.  An  Illinois  bituminous  coal  has  the  following  composition:  C,  62  per 
cent.;  H,  5  per  cent.;  O,  15  per  cent.     Find  the  heat  units  in  the  coal  and 
the  air  required  to  burn  1  Ib. 

5.  A  coking  coal  has  the  following  composition:  C,  85  per  cent.;  H,  5 
per  cent. ;  O,  4  per  cent.     Find  the  heat  value  of  the  coal  and  the  air  required 
to  burn  1  Ib. 

^  6.  A  coal  contains  C,  80  per  cent.;  H,  2  per  cent.;  O,  6  per  cent.  What 
is  its  heat  value  and  how  many  pounds  of  air  will  be  required  to  burn 
1  Ib.  of  it? 

7.  A  coal  contains  C,  70  per  cent.;  H,  5 per  cent.;  O,  8  per  cent.     What 
is  its  heat  value  and  how  much  air  will  be  required  to  burn  1  Ib.  of  it? 

8.  A  coal  has  the  following  composition:  C,  80  per  cent.;  H,  3  per  cent.; 
O,  4  per  cent.     How  much  heat  will  be  lost  if  one-half  of  the  carbon  is  burned 
to  CO  and  the  balance  to  CO2,  and  what  is  the  weight  of  air  required  to  burn 
1  Ib.  of  the  coal  under  these  conditions? 

•~"  9.  A  coal  contains  C,  90  per  cent.;  H,  1  per  cent.;  O,  2  per  cent.  If  three- 
quarters  of  the  carbon  is  burnt  to  CO2  and  the  balance  to  CO,  what  will  be 
the  B.T.U.  given  off  per  pound,  and  what  will  be  the  air  required  to  burn 
1  Ib.  under  the  above  conditions? 

10.  A  flue -gas  shows  the  following  composition:  CO2,  8  per  cent.;  CO, 
0  per  cent. ;  O,  14  per  cent. ;  N,  78  per  cent.     Find  the  pounds  of  air  used 
per  pound  of  coal  if  the  coal  contains  C,  80  per  cent.;  H,  5  per  cent.;  O,  3  per 
cent.;  and  N,  1  per  cent.  ^/.  5  < 

11.  A  flue  gas  shows  the  following  composition:  CO2,  8.1  per  cent.;  CO, 
0  per  cent.;  O,  flrtper  cent.;  Nj^fc&pjer  cent.     Find  the  pounds  of  air  us^d 
per  pound  of  coal,  n  the  coal  contains ^u;  75  per  cent.;  H,  Skper  cent.;  O,  '8  per 
cent.  * 

12.  A  flue  gas  shows  the  following  composition:  CO2,  5  per  cent.;  CO, 
0  per  cent.;  O,  15  per  cent.;  N,  80  per  cent.     Find  the  pounds  of  air  used 
per  pound  of  coal  if  the  coal  contains  C,  75  per  cent.;  H,  5  per  cent.;  O,  8  per 
cent. 

13.  A  flue  gas  shows  the  following  composition:  CO2,  4.1  per  cent.;  CO, 
0  per  cent.;  O,  16  per  cent.;  N,  79.9  per  cent.     Find  the  pounds  of  air  used 
per  pound  of  coal  if  the  coal  contains  C,  75  per  cent.;  H,  5  per  cent.;  O,  8  per 
cent. 

14.  A  flue  gas  shows  the  following  composition:  CO2,  4.3  per  cent.;  CO, 
0  pelf  cent. ;  O,  12.7  per  cent. ;  N,  83  per  cent.     Find  the  pounds  of  air  required 
per  pound  of  coal  if  the  coal  contains  C,  75  per  cent.;  H,  5  per  cent.;  O, 
8  per  cent. 

16.  A  flue  gas  shows  the  following  composition:  CO2,  8.3  per  cent.;  O,  10.8 
per  cent.;  N,  80.9  per  cent.  How  much  air  is  burned  per  pound  of  coal  if  the 
coal  Contains  C,  75  per  cent.;  H,  6  per  cent.;  O,  4  per  cent.? 

16.  A  coal  contains  C,  80 per  cent.;  H,  5  per  cent.;  O,  3 per  cent.;  N,  1  per 
cent.  Find  the  theoretical  temperature  of  combustion  if  30  per  cent,  more 
air  is  used  in  the  combustion  than  is  necessary.  Temperature  of  boiler  room, 
70°. 


COMBUSTION  AND  FUELS  81 

17.  A  coal  has  C,  80  per  cent.;  H,  5  per  cent.;  O,  3  per  cent.;  and  N,  1  per 
cent.     Find  the  theoretical  temperature  of  combustion  if  50  per  cent,  more 
air  is  used  than  is  necessary  for  the  combustion.     Temperature  of  boiler 
rjaom,  80°. 

18.  A  coal  gives  the  following  analysis:  C,  75  per  cent.;  H,  6  per  cent.; 
O,  4  per  cent.;  and  N,  2  per  cent.     Seventy-five  per  cent,  excess  of  air  is  used 
in  burning  it.     What  is  the  ideal  rise  in  temperature  of  the  gases? 


CHAPTER  VI 
BOILERS 

56.  Boilers  may  be  divided,  from  the  path  taken  by  the  fire, 
into  fire-tube  or  tubular  boilers  and  water-tube  or  tubulous  boilers. 
In  the  fire-tube  boiler  the  hot  gases  from  the  fire  pass  through 
the  tubes,  while  in  the  water-tube  boiler  these  gases  pass  around 
the  tubes. 

Boilers  are  also  divided  into  two  classes  depending  on  the 
position  of  the  fire;  these  are  known  as  externally  fired  and  inter- 
nally fired  boilers. 

In  the  externally  fired  boiler,  the  fire  is  entirely  external  to 
the  boiler  and  is  usually  confined  in  a  brick  chamber.  These 
boilers  are  largely  used  for  stationary  plants. 

The  internally  fired  boiler  is  most  commonly  used  for  loco- 
motive and  marine  boilers.  The  fire  is  entirely  enclosed  in  the 
steel  shell  of  the  boiler  and  no  brick  setting  is  necessary.  These 
boilers  are  more  expensive  per  horse-power  than  the  ordinary 
forms  of  stationary  boilers. 

The  various  forms  of  boilers  under  proper  operating  condi- 
tions give  essentially  the  same  economical  results. 

57.  Return  Tubular  Boilers. — Fig.  13  shows  the  plan  and 
elevation  of  the  setting  of  a  fire-tube  boiler  of  the  return  type. 
The  coal  burns  upon  the  grates,  which  rest  upon  the  front  of  the 
boiler  setting  and  upon  the  bridge  wall.  The  flames  pass  under 
and  along  the  boiler  shell,  then  turn  in  the  back  combustion 
chamber  D  and  pass  through  the  tubes  of  the  boiler,  then  out 
through  the  smoke  nozzle  N  and  through  the  breeching  to  the 
chimney.  The  smoke  nozzle  is  shown  at  the  front  of  the  boiler 
setting. 

There  are  usually  two  man-holes  in  the  boiler,  one  in  front 
under  the  tubes  and  one  in  the  top  of  the  boiler.  These  open- 
ings are  reenforced  with  flanged  steel  reinforcements.  The  shells 
are  made  of  boiler  steel  having  a  tensile  strength  of  55,000  to 
66,000  Ibs.  The  shell  of  the  boiler  is  rolled  to  form  and  riveted 
together.  The  heads  of  the  boiler  which  form  the  tube  sheet 

82 


BOILERS 


83 


and  into  which  the  tubes  are  fastened  are  made  of  flanged  steel 
of  about  55,000  Ibs.  tensile  strength.  The  tubes  are  made  of 
steel,  usually  lap  welded.  Charcoal  iron  tubes  are  the  best,  but 
are  difficult  to  get,  so  that  most  manufacturers  use  a  hot-rolled, 
lap-welded  steel  tube. 


Feed  Pipe  M 


Manhole 


1 


1    i 

J  

i 

F 

u 

I 

_-i 

:; 

1 

U 

> 

'5              ! 

•;; 
J> 

—  - 

\ 

3 

J 

1 
1           '^ 

3} 

'      i; 

pi 

i 

; 

1= 

=^ 

N      r 


J 


These  boilers  are  set  in  brick  settings,  and  in  all  brick-set 
boilers  great  care  should  be  taken  in  building  the  setting.  Air 
leaks  in  the  brick  work  should  be  carefully  avoided  as  they  cause 
serious  loss  in  economy.  All  brick  should  be  set  with  full  flush 


84 


HEAT  ENGINES 


mortar  joints  so  as  to  make  the  setting  strong  and  avoid  leakage. 
Fig.  13  shows  the  return  flue  boiler  with  the  boiler  resting  upon 
the  brick  work. 


FIG.  14. — Steel  frame  boiler  support. 


FIG.  15. — Return  tubular  boiler  with  loops  for  suspension  setting. 

Boilers  of  this  type  are  often  supported  by  a  steel  frame- 
work as  shown  in  Figs..  14  and  15.  This  method  is  preferable  as  it 
leaves  the  boiler  independent  of  the  setting.  The  brick  setting  of 


BOILERS 


85 


a  boiler  has  very  little  strength  and  this  arrangement  leaves  the 
boiler  setting  free  from  all  strain  due  to  the  weight  of  the  boiler. 
In  earlier  boiler  construction  it  was  customary  to  place  a 
steam  dome  on  all  boilers.     The  object  of  doing  this  was  to 


FIG.   16. — Dry  pipe. 

provide  dry  steam.  Most  engineers  have  discarded  the  use  of 
steam  domes  on  high-pressure  boilers  as  they  weaken  the  boiler 
shell  and  add  to  the  expense  of  the  boiler  construction.  To 
avoid'  getting  wet  steam  from  the  boiler  a  dry-pipe  is  provided 
as  shown  in  Fig.  16. 


FIG.  17. — Brick  setting  for  fire-tube  boiler  with  overhanging  shell. 

Fig.  17  shows  a  return  flue  boiler  and  solid  brick  setting. 
Some  engineers  prefer  a  setting  having  a  2-in.  air  space  in  the 
center  of  the  wall.  The  brick  walls  enclosing  a  fire-tube  boiler 
are  made  very  heavy  so  as  to  give  good  heat  insulation,  preventing 
anexcessive  loss  of  heat  from  the  boiler,  and  also  to  prevent  the 


86 


HEAT  ENGINES 


BOILERS 


87 


filtration  of  air  through  the  setting  and  the  consequent  cooling 
of  the  hot  gases  passing  away  from  the  fire. 

58.  Internally  Fired  Boilers. — Another  large  class  of  return 
tubular  boilers  are  the  internally  fired  boilers.  These  boilers 
have  been  extensively  used  for  marine  purposes.  Fig.  18  shows 
an  internally  fired  Scotch  marine  boiler.  The  cut  shows  two 
internal  furnaces.  In  the  larger  sizes  these  boilers  are  often 


FIG.  19. — Scotch  marine  boiler. 

made  with  three  or  even  four  furnaces  (see  Fig.  19).  These  can 
be  built  in  large  sizes,  and  are  very  compact,  making  them  par- 
ticularly suitable  for  marine  work. 

Fig.  20  shows  one  of  these  boilers  built  for  stationary  pur- 
poses. The  steel  back  combustion  chamber  used  in  marine  work, 
shown  in  Fig.  18,  is  replaced  by  brick  construction  in  Fig.  20. 
In  very  large  boilers  of  this  type,  furnaces  are  provided  at  each 
end,  opening  into  a  common  combustion  chamber  in  the  middle 
of  the  boiler. 


88 


HEAT  ENGINES 


BOILERS 


89 


59.  Locomotive  Type  of  Boiler. — A  special  type  of  fire-tube 
boiler  is  used  on  locomotives.  In  this  boiler  the  combustion 
space,  including  the  grates,  and  the  sides  of  the  ash  pit  are  sur- 
rounded by  a  water  space.  The  gases  pass  directly  from  the 
fire  through  the  tubes  and  up  the  stack.  As  in  the  internally 
fired  boiler,  the  hot  gases  do  not  come  in  contact  with  the  shell 
of  the  boiler.  This  permits  of  the  use  of  higher  pressures  in  these 
boilers,  often  as  high  as  225  Ibs.  Modifications  of  this  type  of 
boiler  are  used  for  threshing  and  other  types  of  portable  boilers. 
They  are  sometimes  used  for  stationary  purposes,  particularly 


FIG.  21. — Locomotive  type  of  boiler. 

for  heating  where  a  compact  form  of  boiler  is  desirable.  Fig. 
21  shows  the  side  elevation  of  a  boiler  of  this  class  designed  for 
stationary  use. 

60.  Use  of  Tubular  Boilers. — The  fire-tube  boiler,  as  shown 
in  Fig.  13,  has  certain  limitations  in  use.  Its  construction  is  such 
that  hot  gases  pass  outside  the  shell,  with  cold  water  on  the  inside 
of  the  shell.  This  produces  a  large  difference  of  temperature  on 
the  two  sides  of  the  shell,  and  a  strain  is  produced  in  the  metal 
of  the  shell,  owing  to  this  difference  of  temperature.  The  thicker 
the  shell  the  greater  is  the  difference  in  temperature  between  the 
two  sides  of  the  shell.  In  practice  it  is  found  that  the  thickness 
of  the  shell  should  not  exceed  \  in.  This  limitation  in  the 
thickness  of  the  shell  limits  the  diameter  of  the  boiler  and  the 
pressure  that  the  boiler  can  carry.  It  is  customary  to  use  this 


90 


HEAT  ENGINES 


class  of  boilers  for  pressures  not  to  exceed  125  Ibs.  per  square 
inch  and  in  sizes  not  larger  than  125  boiler  horse-power. 

A  majority  of  the  more  recent  plants  are  being  operated  at 
over  125  Ibs.  pressure  and  therefore  a  fire-tube  boiler  cannot  be 
used.  In  addition  the  horse-power  of  each  boiler  unit  is  so  small 
that  a  very  large  number  of  boiler  units  would  be  necessary.  In 
a  power  plant  of  say  50,000  horse-power,  such  as  exists  in  the 
larger  cities,  if  this  type  of  boiler  were  used,  there  would  be 
required  400  boilers  and  the  space  required  for  this  number  of 
units  would  make  it  almost  impossible  to  install  such  a  plant. 


Safety  Valve 


FIG.  22. — Babcock  and  Wilcox  boiler. 

The  internally  fired  boiler  is  not  as  limited  in  the  pressure 
that  it  can  carry  as  is  the  return  fire-tube  type,  since  the  fire  does 
not  come  in  contact  with  the  boiler  shell  and  the  shell  can  be 
made  thicker.  The  increased  thickness  of  shell  permits  the 
building  of  larger  boilers  of  this  type  than  of  the  return  fire  tube, 
and  they  have  been  built  in  units  of  500  horse-power  carrying 
200  Ibs.  pressure.  They  have  not  been  much  used  for  stationary 
purposes  owing  to  their  first  cost  and  the  cost  of  repairs  where 
conditions  are  not  favorable  to  their  use. 


BOILERS 


91 


Hand  Hole 

/  PlatesN 


61.  Water-tube  Boilers. — The  demand  for  increased  pressure 
and  for  larger  sized  boiler  units  has  led  to  the  introduction  of 
water-tube  boilers,  and  all  the  larger  power  stations  to-day  are 
using  water-tube  boilers  almost  exclusively.  The  principal 
reasons  for  using  the  water-tube  boilers  in  large  power  stations 
are:  adaptability  to  high  pressure,  reduced  space  taken  by  the 
boiler,  and  greater  safety  in  operation.  There  are  a  great  many 
different  makes  of  water-tube 
boilers  on  the  market  of  various 
types,  both  vertical  and  horizontal. 

Fig.  22  shows  a  Babcock  and 
Wilcox  boiler  in  longitudinal  cross- 
section.  Gases  from  the  fire  pass 
up  through  the  tubes,  being  de- 
flected vertically  by  a  baffle  wall 
located  between  the  tubes  and 
directly  above  the  bridge  wall. 
They  then  pass  down  around  the 
tubes  to  the  space  back  of  the 
bridge  wall,  being  deflected  by 
another  baffle,  then  up  between 
the  tubes  and  out  through  the 
smoke  opening  which  is  in  the  rear 
of  the  boiler  setting  and  above  the 
tubes. 

As  it  is  heated,  the  water  in  the 
tubes  tends  to  rise  toward  their  up- 
per, or  front  end,  then  rises  through 

the  front  header  and  connection  into  the  steam  and  water  drum, 
where  the  steam  separates  from  the  water,  and  the  latter  flows 
back  in  the  drum  and  down  through  the  rear  header.  The  feed 
water  enters  the  boiler  through  a  pipe  passing  through  the  front 
end  of  the  drum  and  extending  back  about  one  third  its  length. 

Fig.  52  shows  a  Babcock  and  Wilcox  boiler  with  a  superheater 
attached. 

Fig.  23  shows  the  front  and  side  views  of  a  header  in  a  Babcock 
and  Wilcox  boiler  and  indicates  clearly  the  way  the  tubes  are 
"staggered." 

This  class  of  boiler  gives  very  satisfactory  service  for  high- 
pressure  work,  having  large  disengaging  surfaces  for  the  steam  to 
leave  the  water,  and  ample  ste'am  space. 


Side  view  of 
vertical  header 


Front  view  of 
vertical  header 


FIG.  23. — Tube   header  in  Bab- 
cock and  Wilcox  boiler. 


92 


HEAT  ENGINES 


Fig.  24  shows  a  sectional  side  elevation  of  a  Stirling  boiler. 

"This  consists  of  three  transverse  steam  and  water  drums 
set  parallel  and  connected  to  one  mud  drum  by  water  tubes  so 
curved  that  their  ends  enter  the  tube  sheets  at  right  angles  to 
the  surface.  This  curvature  of  the  tubes  gives  ample  and  effi- 
cient provision  for  expansion  and  contraction.  The  front  and 
middle  steam  drums  are  connected  by  curved  ^equalizing  tubes 


FIG.  24. — Stirling  boiler. 

above  the  water  line  and  curved  circulating  tubes  below  the 
water  line,  while  the  rear  and  middle  drums  are  connected  by 
curved  equalizing  tubes  above  the  water  line  only. 

The  steam  generated  in  the  three  banks  of  tubes  passes  into 
the  middle  drum,  which  is  set  higher  than  the  other  two  to 
give  additional  steam  space,  thence  it  passes  through  the  main 
steam  outlet,  which  may  be  located  anywhere  along  the  top  of 
the  drum. 


BOILERS  93 

The  safety  valves  are  located  on  the  top  of  the  middle  steam 
drum.  The  feed  pipe  connection  passes  through  the  top 
of  the  rear  drum  into  a  trough  by  which  the  water  is  dis- 
tributed along  the  whole  length  of  the  drum.  The  blow- 
off  connection  is  attached  to  the  bottom  of  the  mud  drum  at 
the  center  and  passes  out  through  a  sleeve  in  the  rear  wall, 
just  outside  of  which  the  blow-off  valve  is  located.  The  water 
column,  located  at  one  side  of  the  front  of  the  boiler,  is  con- 
nected to  one  head  of  the  center  steam  and  water  drum.  The 
feed  water  enters  the  upper  rear  drum  and  passes  downward 
through  the  rear  bank  of  tubes  to  the  lower  drum,  thence  up- 
ward through  the  front  bank  to  the  forward  steam  and  water 
drum.  The  steam  formed  during  the  passage  upward  through 
the  front  bank  of  tubes  becomes  separated  from  the  water  in 
the  front  drum,  and  passes  through  the  upper  row  of  cross 
tubes  into  the  middle  drum,  from  which  point  it  enters  the 
steam  main.  The  water  from  the  front  drum  passes  through 
the  lower  cross  tubes  into  the  middle  drum,  and  thence  down- 
ward through  the  middle  bank  of  tubes  to  the  lower  drum, 
from  which  it  is  again  drawn  up  the  front  bank  to  retrace  its 
former  course.  The  steam  generated  in  the  rear  bank  of  tubes 
passes  through  the  cross  tubes  to  the  center  drum.  In  its  pas- 
sage down  the  rear  bank  of  tubes  the  feed  water  is  heated  so 
that  much  of  the  scale-forming  matter  is  precipitated  and 
gathers  in  the  rear  bank  of  tubes  and  in  the  mud  drum,  where 
it  is  protected  from  high  temperatures  and  can  be  washed  and 
blown  out  as  frequently  as  tire  case  demands." 

The  hot  gases  circulate  in  the  reverse  direction.  On  leaving 
the  fire  they  are  deflected  by  baffle  walls  so  as  to  pass  up  between 
the  tubes  to  the  first  drum,  then  down  around  the  tubes  from  the 
second  drum,  and  again  up  between  the  tubes  to  the  rear  drum. 
The  burned  gases  leave  the  boiler  at  the  rear  near  the  upper  end 
of  the  last  bank  of  tubes.  This  boiler  represents  the  ideal  cir- 
culation as  far  as  the  paths  of  the  water  and  gases  are  concerned; 
that  is,  the  coldest  gases  come  in  contact  with  the  coldest  water  in 
the  boiler,  and  the  hottest  gases  come  in  contact  with  the  hottest 
water.  The  drums  with  their  connecting  tubes  are  supported 
by  a  steel  frame  built  into  the  brick  work  of  the  boiler.  The  brick 
setting  only  serves  to  enclose  the  gases  and  is  under  no  strain 
due  to  the  weight  of  the  boiler. 


94  HEAT  ENGINES 

There  is  a  man-hole  in  one  end  of  each  of  the  four  drums  and  by 
the  removal  of  the  man-hole  plates  the  drums  may  be  entered. 

Fig.  25  shows  a  cross-section  of  a  Heine  water-tube  boiler. 
In  this  boiler  the  gases  of  combustion  pass  over  the  bridge  wall 
into  the  combustion  chamber,  where  they  are  completely  burned. 
They  then  pass  upward  back  of  the  lower  baffle  wall  (which 
consists  of  a  row  of  tiling)  and  then  forward  around  the  tubes, 
and  parallel  to  them,  to  the  front  of  the  boiler,  where  they  turn 
up  in  front  of  the  forward  end  of  the  upper  baffle  wall  and  then 


FIG.  25. — Heine  boiler. 

pass  back  around  the  shell  to  the  opening  to  the  breeching. 
The  feed  water  enters  the  boiler  through  the  front  head,  pass- 
ing into  the  mud  drum  where  the  dirt  and  sediment  are  de- 
posited, then  flows  back  along  the  bottom  of  the  drum  and 
then  forward  along  the  top  and  out  of  the  drum  at  the  front 
end.  From  here  the  circulation  is  toward  the  back  of  the  boiler, 
down  the  rear  water-leg,  forward  through  the  tubes,  and  up  the 
front  water-leg  into  the  boiler  again.  The  steam,  which  is  formed 
very  largely  in  the  tubes,  is  carried  along  with  the  water  and 
discharged  into  the  boiler  from  the  front  water-leg. 

Fig.  26  shows  a  side  elevation  of  the  water-legs,  shell  and 
tubes  in  the  Heine  boiler. 


BOILERS 


95 


Where  a  plant  is  very  limited  in  the  floor  space  available,  it 
is  often  desirable  to  use  a  vertical  water-tube  boiler.  Fig.  27 
shows  a  cross-section  of  the  Wickes  vertical  boiler.  The  grates 
are  located  in  a  " Dutch  Oven"  front  built  out  from  the  main 
boiler  setting.  The  gases  pass  up  around  the  tubes  in  the  for- 
ward half  of  the  boiler  and  down  around  them  in  the  rear  half, 
leaving  the  boiler  in  the  rear  near  the  lower  drum.  The  water 
inside  the  tubes  flows  in  the  same  direction  as  the  gases,  in  both 
the  front  and  rear  compartments.  These  boilers  are  quick 
steamers  and  occupy  relatively  small  floor  space. 

Fig.  28  shows  a  vertical  boiler  of  the  Rust  type.  The  furnace 
and  combustion  chamber  project  from  the  front  of  the  boiler 
as  a  " Dutch  Oven"  front. 


FIG.  26. — Heine  boiler  showing  water-legs,  shell  and  tubes. 

"  The  products  of  combustion  travel  up  the  first  pass,  down 
the  second  pass  and  out  to  the  stack. 

"The  water,  which  is  fed  into  the  center  of  the  rear  water-and- 
mud  drum,  passes  across  through  the  circulating  tubes  to  the 
front  water-and-mud  drum,  up  the  vertical  tubes  in  the  front 
pass  to  the  front  steam-and-water  drum  where  the  steam 
generated  in  the  front  pass  is  separated  from  the  water.  The 
water  then  passes  over  through  the  circulating  tubes  to  the  rear 
steam-and-water  drum  and  down  the  rear  bank  of  tubes  to  the 
starting  place. 

"The  steam  liberated  in  the  front  steam-and-water  drum  passes 
through  the  steam  tubes  into  the  steam  space  of  the  rear  steam- 
and-water  drum,  the  entrained  moisture  dropping  into  the 


96 


HEAT  ENGINES 


Steam  Outlet 


Manhole 


Safety  Valve 


Water 
Column 


Feed  Inlet 


-  Downcomers1 


FIG.  27. — Wickes  boiler. 


BOILERS 


97 


water  space,  while  the  steam  passes  along  the  top  of  the  drum 
through  the  dry  pipe  to  the  steam  outlet." 

62.  Horse-power  Rating  of  Boilers. — The  term  "  horse- 
power," as  applied  to  boilers,  has  no  definite  value  and  is  only 
used  as  a  matter  of  convenience.  The  ability  of  a  boiler  to 

,  Steam  and  Water  Drums  -, 

Steam  \ 

\    Steam  ;      Tubes  ! 

\   Outlet 


Water 
Circulating 
Tubes-- 


Water 
Column 


^-Blow-off  ' 
' *  Water  and  Mud  Drums-'' 
FIG.  28.— Rust  boilef. 

make  steam  depends  on  the  amount  of  heating  surface  in  it. 
Experience  has  determined  that  for  the  best  results  in  the 
ordinary  form  of  boiler,  a  square  foot  of  heating  surface  should 
not  evaporate  more  than  3  Ibs.  of  water  per  hour  (if  economy 

7 


98  HEAT  ENGINES 

is  highly  desired).  In  writing  specifications  for  boilers, 
it  is  customary  to  state  the  number  of  square  feet  of  heating 
surface  the  boiler  is  to  contain  and  the  pounds  of  water  it  is  to 
evaporate  per  hour  under  the  given  conditions,  rather  than  the 
boiler  horse-power.  In  order  to  give  the  term  "  boiler  horse- 
power" a  definite  meaning,  the  American  Society  of  Mechanical 
Engineers  has  adopted  the  following  rating  for  boilers:  A 
"boiler  horse-power"  is  34.5  Ibs.  of  water  evaporated  per  hour  from 
and  at  212°  into  dry  and  saturated  steam.  Most  boilers  will 
produce  from  25  to  50  per  cent,  more  steam  than  their  rating, 
depending  upon  the  amount  of  heat  generated  in  the  furnace 
and  the  amount  of  heat  that  is  given  to  the  water  in  the  boiler. 
The  amount  of  heat  given  off  by  the  fuel  will  depend  upon  the 
kind  of  fuel  used,  the  area  of  the  grate,  the  amount  of  draft, 
and  the  skill  of  the  fireman.  A  very  rapid  rate  of  combustion 
usually  results  in  a  large  escape  of  heat  to  the  stack  and  reduced 
economy. 

There  is  no  relation  between  a  boiler  horse-power  and  an  engine 
horse-power.  The  number  of  boiler  horse-power  required  to 
supply  steam  for  a  given  engine  horse-power  will  be  determined 
by  the  number  of  pounds  of  steam  the  engine  requires  to  develop 
a  horse-power.  The  steam  required  per  horse-power  hour  varies 
through  a  wide  range  in  the  different  types  of  engines. 

63.  Heating  Surface,  Grate  Surface,  and  Breeching. — The 
water  heating  surface  in  a  boiler  is  that  part  of  the  boiler  which 
has  water  on  one  side  and  hot  gases  on  the  other.  Superheating 
surface  has  steam  on  one  side  and  hot  (jases  on  the  other.  In  both 
cases  the  side  in  contact  with  the  hot  gases  is  the  one  to  be  measured. 
The  proportion  of  grate  surface  to  heating  surface  depends  upon 
the  kind  of  fuel  and  the  intensity  of  the  draft.  In  small  boilers 
such  as  are  used  for  heating  purposes,  with  light  draft  and  hard 
coal  it  is  usual  to  allow  1  sq.  ft.  of  grate  to  from  20  to  30*  sq. 
ft.  of  heating  surface.  In  large  power  boilers  the  ratio  of  grate 
surface  to  heating  surface  varies  from  1  to  50,  to  from  1  to  70. 
In  locomotive  boilers  with  forced  draft  the  ratio  is  from  1  to 
50,  to  1  to  100. 

The  rate  of  combustion  varies  with  the  kind  of  coal  and  with 
the  draft.  With  anthracite  coal  and  moderate  draft,  not  ex- 
ceeding five-tenths  of  an  inch  of  water,  it  is  from  12  to  15  Ibs. 
per  square  foot  of  grate  surface  per  hour,  and  with  bituminous 
coal  from  15  to  20  Ibs.  The  air  opening  in  the  grate  depends 


BOILERS 


99 


upon  the  kind  of  coal  and  usually  does  not  exceed  50  per  cent. 
of  the  grate  area.  Anthracite  and  the  better  grades  of  bitu- 
minous coal  require  less  air  opening  than  the  poorer  grades  of 
coal. 

The  following  rule  is  used  for  determining  the  heating  surface 
of  a  horizontal  return  flue  fire-tube  boiler:  the  heating  surface  is 
equal  to  two-thirds  the  cylindrical  surface  of  the  shell,  plus  the 
internal  area  of  all  the  tubes,  plus  two-thirds  the  area  of  both  tube 
sheets,  minus  twice  the  combined  external  cross-sectional  area  of 
all  the  tubes,  all  expressed  in  square  feet. 

In  water-tube  boilers  it  is  customary  to  allow  10  s_q.  ft.  of 
heating  surface  per  bQi]grrsp-pnwr,nrl  in  fire-tube  boilers 


12  sq.  ft. 

The  connection  for  carrying  the  hot  gases  from  the  boiler 
to  the  chimney  is  called  the  breeching.  The  area  of  the  breech- 
ing is  from  J  to  |  of  the  area  of  the  grates,  depending  on  the 
strength  of  the  draft.  The  breeching  is  usually  made  of  sheet 
steel  well  braced,  and  should  be  provided  with  a  door  for  cleaning 
and  inspection. 

TABLE  XVI.  DIAMETER  OF  BOILER  TUBES 


Outside 


Inside 


Inches 

Feet                                  Inches 

Feet 

2 

.167 

1.80 

.150 

2£ 

.208 

2.28 

.190 

3 

.250 

2.78 

.232 

3£ 

.292 

3.26 

.272 

4 

.333 

3.74 

.312 

4£               .            .375 

4.24 

.353 

5                              .417 

4.72 

.393 

64.  Boiler  Economy. — The  economy  of  a  boiler  is  usually  ex- 
pressed as  the  number  of  pounds  of  water  fed  to  the  boiler  per 
pound  of  coal  fired. 

The  water  evaporated  by  a  boiler  is  equal  to  the  weight  of 
water  fed  to  the  boiler  (corrected  for  leakage),  provided  that  the 
steam  formed  is  dry  and  saturated.  If  the  steam  is  wet,  it  is 
necessary  to  make  a  correction  in  order  to  determine  how  much 
dry  and  saturated  steam  might  have  been  formed.  When  the 
percentage  of  moisture  is  less  than  2  per  cent,  this  correction 
may  be  made  by  simply  subtracting  the  moisture  from  the  total 
weight  of  water  fed.  If  the  percentage  is  more  than  2  per  cent. 


100  HEAT  ENGINES 

or  if  great  accuracy  is  desired,  the  weight  of  water  fed  must  be 
multiplied  by  a  " factor  of  correction"  which  is  equal  to 

/  h  -  (t  -  32)   \ 
ff  +  (1  -  ff)  {  H^T-  ~d)  } 

where  q  is  the  quality  of  the  steam,  h  is  the  heat  of  the  liquid 
and  H  the  total  heat  of  the  steam  at  the  given  pressure,  and  t 
the  temperature  of  the  feedwater. 

In  order  to  compare  boilers  working  under  different  conditions 
of  feed  temperature  and  steam  pressure  and  with  different  coals,  it 
is  better  to  reduce  them  all  to  the  same  conditions,  and  the 
economy  may  be  expressed  as  the  number  of  pounds  of  equivalent 
evaporation  from  and  at  212°  per  pound  of  combustible  burned. 
By  "equivalent  evaporation  from  and  at  212° "  is  meant  the  number 
of  pounds  of  water  that  would  be  evaporated  from  a  feed  temperature 
of  212°  into  dry  and  saturated  steam  at  212°  by  the  expenditure  of 
the  same  amount  of  heat  as  is  actuall^ased  in  evaporating  the  water 
under  the  given  conditions.  The  "factor  of  evaporation"  is  that 
factor  by  which  the  water  evaporated,  corrected  for  moisture  in  the 
steam,  must  be  multiplied  in  order  to  get  the  equivalent  evaporation. 
It  is  equal  to  the  heat  necessary  to  make  1  Ib.  of  dry  and  satu- 
rated steam  under  the  given  conditions  divided  by  the  heat 
necessary  to  make  1  Ib.  from  and  at  212°. 

With  a  good  boiler  and  high-grade  bituminous  coal,  a  boiler 
will  evaporate  from  9  to  12  Ibs.  of  water  per  pound  of  coal. 
The  average  performance  under  usual  working  conditions  is 
from  8  to  10  Ibs.  of  water  per  pound  of  coal.  The  economy 
of  boiler  operation  depends  not  only  upon  the  construction  of 
the  boiler,  but  also  upon  the  skill  of  the  fireman.  This  is  par- 
ticularly true  with  hand  firing,  and  a  careful  record  of  the  fire- 
man should  be  kept,  in  order  to  prevent  a  waste  of  coal  due  to 
improper  handling  of  the  fires. 

65.  Efficiency  of  Steam  Boilers. — The  efficiency  of  boiler, 
furnace  and  grate  is  the  ratio  of  the  heat  absorbed  per  pound  of 
dry  coal  fired,  to  the  heating  value  of  a  pound  of  dry  coal. 

The  efficiency  of  boiler  and  furnace  is  the  ratio  of  the  heat 
absorbed  per  pound  of  combustible  burned  to  the  heating  value 
of  a  pound  of  combustible. 

The  "heat  absorbed"  per  pound  of  dry  coal  or  combustible  is 
found  by  multiplying  the  equivalent  evaporation  from  and  at 
212°  F.  per  pound  of  dry  coal,  or  combustible,  by  970.4. 


BOILERS 


101 


The  "dry  coal  fired"  is  found  by  deducting  the  moisture  in 
the  coal  from  the  total  weight  of  coal  supplied  to  the  grates. 

The  "combustible  burned"  is  determined  by  deducting  from 
the  weight  of  coal  fired,  the  weight  of  moisture  in  the  coal  plus 
the  weight  of  ash  and  refuse  taken  from  the  ash  pit  plus  "the 
weight  of  dust  and  soot,  if  any,  withdrawn  fron  the  tubes,  flues 
and  combustion  chamber,  including  ash  carried  away  in  the 
gases,  if  any,  determined  from  the  analysis  of  coal  and  ash." 

The  "heating  value  of  a  pound  of  combustible"  is  equal  to 
the  heating  value  of  a  pound  of  dry  coal  divided  by  1  minus  the 
percentage  of  ash  in  the  dry  coal  as  shown  by  analysis. 

Actual  tests  of  various  boilers  show  that  the  efficiency  under 
ordinary  working  conditions  varies  from  60  to  80  per  cent. 
Seventy  per  cent,  might  be  considered  as  a  good  average  efficiency. 

66.  Losses  in  Boiler. — The  principal  losses  in  a  boiler  are  the 
heat  that  is  carried  away  by  the  flue  gases,  the  loss  due  to 
hydrogen  in  the  coal,  the  loss  through  the  grates,  and  the  loss 
by  radiation.  Of  these,  the  largest  is  the  heat  carried  up  the 
chimney  by  the  stack  gases.  The  following  table  shows  the 
relative  proportions  of  these  losses  in  a  well-operated  boiler 
plant,  and  is  termed  the  heat  balance.  The  total  heat  in  1 
Ib.  of  combustible  in  the  coal  was  15,070  B.T.U. 

TABLE  XVII.  HEAT  BALANCE  IN  BOILER  PLANT 


Distribution  of  heat  of  dry  coal 

B.T.U. 

Per  cent. 

1    Heat  absorbed  by  the  boiler 

10982 

77  82 

2.  Loss  due  to  evaporation  of  moisture  in  the  coal  .  . 
3.  Loss  due  to  heat  carried  away  by  steam  formed 
by  the  burning  of  hydrogen  in  the  coal 
4.  Loss  due  to  heat  carried  away  in  dry  chimney 
gases. 
5.  Loss  due  to  carbon  monoxide     

27 
628 

1,635 
96 

.19 
4.45 

11.57 
68 

6    Loss  due  to  combustible  in  ash  and  refuse 

319 

2  26 

7.  Loss  due  to  heating  moisture  in  the  air  

40 

.28 

8.  Loss  due  to  unconsumed  hydrogen  and  hydro- 
carbons, to  radiation,  and  unaccounted  for. 

388 

2.75 

Total 

14,115 

100  0 

The  heat  carried  away  by  the  chimney  gases  depends  upon 
the  amount  of  air  admitted  to  the  fire  and  upon  the  tempera- 
ture at  which  the  gases  leave  the  boiler.  In  a  properly  operated 
plant,  the  gross  loss  of  heat  up  the  chimney  should  not  exceed 


102  HEAT  ENGINES 

20  per  cent.  It  is  often  much  more  than  this  owing  to  the  fact 
that  the  fireman  admits  too  much  air  to  the  coal;  more  than 
is  necessary  for  its  complete  combustion.  This  excess  of  air 
is  heated  from  the  temperature  of  the  boiler  room  to  the  tem- 
perature of  the  stack  gases,  and  all  the  heat  used  for  this  pur- 
pose passes  up  the  chimney  and  is  wasted.  It  is,  therefore,  very 
important  that  the  amount  of  air  admitted  to  the  fire  should 
not  be  more  than  is  absolutely  necessary.  This  is  determined 
by  the  amount  of  carbon  dioxide  in  the  stack  gas  analysis  which 
has  been  previously  described.  In  a  well-operated  plant,  the 
CO2  as  shown  by  the  analysis  varies  from  9  to  10  per  cent.,  and 
under  exceptional  conditions  an  analysis  showing  16.8  per  cent, 
of  C02  has  been  obtained.  It  is  usually  undesirable  to  have  more 
than  12  to  13  per  cent,  of  CO2  in  the  stack  gases.  Larger  per- 
centages generally  indicate  the  presence  of  CO. 

Example.— A  48  in.'  X  12  ft.  return  flue  fire-tube  boiler  has  thirty  4-in. 
tubes.  It  evaporates  1400  Ibs.  of  water  per  hour  from  a  feed  temperature 
of  120°  into  steam  at  100  Ibs.  What  per  cent,  of  its  rating  is  the  boiler 
developing? 

Solution. — First  find  heating  surface  from  the  rule  in  paragraph  63. 
H.S.  of  cylindrical  portion  of  shell  ^-*»  , 

=     ?~X-3.l4i6  X  4  X  12.        =100. 6  sq.ft. 

—        3^74 

H.S.  of  tubes  =  30  X  3.1416  X  -^r*  X  12  =  352.5  sq.  ft. 

-.  •  •  \.£i  -   -    ' 

H.S.  of  tube  sheets 

=  2[f(3.1416  XJ2  X  2)  -(30  X  3.1416  X  i  X  « 

=  2(8 . 38  -  2 . 62)  =11 . j>  sq.  ft. 

Total  heating  surface  =  464 . 5  sq.  ft. 

From  paragraph  63,  the  rate4  horse-power  =         '     =  38 . 7 

Now  find  actual  horse-power  developed. 

The  heat  actually  used  in  evaporating  a  pound  of  water  is  equal  to  the 
total  heat  in  a  pound  of  steam  at  thfc  given  pressure  minus  the  heat 
already  in  the  feed  water. 

The  heat  used  in  evaporating  water  under  actual  conditions 

=  1400  [1188. 6  -  (120  -  32)]  =  1,541,000  B.T.U. 

From  paragraph  64,  the  equivalent  evaporation  from  and  at  212° 

1,541,000 
=  -Q7Q  A      =  1588  Ibs.  per  hour, 

and  from  paragraph  62,  the  boiler  horse-power 


BOILERS  103 

1588 


34.5 


46. 


46 
5^  =  1.19  =  119  per  cent. 

00.   I 

Ans.     Boiler  is  developing  19  per  cent,  overload. 

Example.  —  A  boiler  evaporates  8.23  Ibs.  of  water  per  pound  of  coal 
fired.  Feed  temperature,  120°;  steam  pressure,  100  Ibs.  Coal  as  fired 
contains  2  per  cent,  moisture.  Dry  coal  contains  5  per  cent,  ash  and 
has  a  heating  value  of  12,800  B.T.U.  per  pound.  Twelve  per  cent,  of 
coal  fired  is  taken  from  ash  pit  in  form  of  ash  and  refuse,  (a)  Find  the 
efficiency  of  the  boiler,  furnace  and  grates  combined.  (6)  Find  the 
efficiency  of  the  boiler  and  furnace. 

Solution.  —  (a)  Heat  necessary  to  evaporate  1  Ib.  of  water 

=  1188.6  -  (120  -  32)  =  1100.6  B.T.U. 
Water  evaporated  per  pound  of  dry  coal  fired 

8.23  8.23 

=  1.  00^02  =T98  = 

Heat  utilized  per  pound  of  dry  coal  fired 

=  8.4  X  1100.6  =  9245  B.T.U. 
Efficiency  of  boiler,  furnace  and  grates  combined 

_  Heat  utilized  per  pound  of  dry  coal  fired 
Heating  value  of  1  Ib.  of  dry  coal 

0045 
=  =  .7223  =  72.23  per  cent. 


(6)  Heating  value  of  1  Ib.  of  combustible 

12800          12800         , 
=  EOO"-  .05  =  ~96~  "  13'474  RT-U' 

Water  evaporated  per  pound  of  combustible  burned 

8'13  =  9.57  Ibs. 


1.00  -(.02+  .12)        .86 
Heat  utilized  per  pound  of  combustible  burned 

=  9.57  X  1100.6  =  10,533  B.T.U. 

Efficiency  of  boiler  and  furnace 

_  Heat  utilized  per  pound  of  combustible  burned    , 
Heating  value  of  1  Ib.  of  combustible 


104  HEAT  ENGINES 

10533 


13474  ~  -7817"  =  78-17  Percent.* 

.        f  (a)  72.23  per  cent. 
?<\(6)  78. 17  per  cent. 


Example.  —  If  26  Ibs.  of  air  are  used  to  burn  a  pound  of  coal  containing 
13,500  B.T.U.,  and  the  temperature  of  the  stack  gases  is  550°,  what  per 
cent,  of  heat  is  lost  up  the  stack,  if  the  temperature  of  the  boiler  room  is 
70°? 

Solution.  —  If  there  were  no  ash  in  the  coal,  each  pound  burned  would 
give  off  a  pound  of  gas  and  the  total  weight  of  stack  gas  per  pound  of  coal 
fired  would  be  26  +  1  =  27  Ibs.  This,  however,  is  never  the  case  as 
there  is  always  some  ash  and  unburned  coal,  and  hence  the  actual  weight 
of  stack  gas  per  pound  of  coal  is  something  a  little  less  than  27  Ibs.  The 
average  of  the  specific  heats  of  the  various  components  of  the  stack  gase? 
is  a  little  higher  than  that  of  air,  .2375.  To  be  absolutely  correct,  then, 
it  would  be  necessary  to  multiply  the  weight  of  each  of  the  various  gases 
in  the  stack  gas  by  its  particular  specific  heat,  and  then  add  these  prod- 
ucts together  to  get  the  B.T.U.  necessary  to  raise  the  products  of  com- 
bustion one  degree.  This,  however,  is  never  done,  the  method  used 
being  to  assume  the  specific  heat  of  the  stack  gases  to  be  the  same  as  that 
of  air,  .2375,  although  really  it  is  slightly  higher,  and  to  assume  that  1 
Ib.  of  gas  is  given  off  from  1  Ib.  of  coal,  although  in  reality  it  is  a  little 
less.  Thus  one  assumption  practically  offsets  the  other,  and  the  result 
is  approximately  correct. 

Hence,  the  heat  necessary  to  raise  the  products  of  combustion  one 
degree 

=  .2375(26  +  1)  =  6.41  B.T.U. 

Rise  in  temperature  of  the  stack  gases 
=  550  -  70  =  480°. 

*  This  answer  may  be  checked  as  follows: 
Efficiency  of  grate  alone 

_  Combustible  burned  per  pound  of  coal  fired 
Combustible  fired  per  pound  of  coal  fired 

=      =  -9237  =  92-37  per  cent- 


Efficiency  of  boiler  and  furnace 

_  Efficiency  of  boiler,  furnace  and  grate 
Efficiency  of  grate  alone 

722S  - 

=  1)237  =  -7819  =78-19Per  cent- 


BOILERS  105 

Heat  necessary  to  raise  the  stack  gases  480° 

=  480  X  6.41  =  3080B.T.U. 
Per  cent,  of  heat  lost  up  the  stack 

3080 
=  1Q™  =.2281  =  22.81  per  cent. 

loOUU 

67.  Boiler  Accessories. — In  order  to  determine  the  physical 
condition  of  the  steam  and  water  in  a  boiler,  all  boilers  are 
provided  with  a  steam  gage  showing  the  pressure  per  square 
inch  in  the  boiler,  a  gage  glass  to  indicate  the  water  level  in 
the  boiler,  and  a  safety  valve  which  automatically  relieves  the 
pressure  in  the  boiler  should  it  exceed  the  safety  point.  The 
feed-water  pump,  or  other  feeding  device,  supplies  the  boiler 


Elevation.  Interior  mechanism 

FIG.  29. — Pressure  gage. 

with  water  to  take  the  place  of  water  which  has  been  made  into 
steam.  The  blow-off  cock  is  attached  to  the  lowest  point  of 
the- boiler  and  drains  the  water  from  the  boiler.  This  is  usually 
opened  from  time  to  time  to  blow  the  mud  and  settlings  out 
of  the  boiler. 

The  ordinary  form  of  pressure  gage  is  shown  in  Fig.  29. 
Pressure  gages  should  be  placed  at  a  convenient  point  for  easy 
observation,  and  the  piping  should  be  as  short  as  possible.  The 
gage  should  always  be  provided  with  a  siphon  containing  water  so 
that  the  hot  steam  cannot  enter  the  gage.  If  hot  steam  enters  the 
gage  it  changes  the  length  of  the  copper  gage-tube,  which  changes 
the  calibration  of  the  instrument.  It  should  also  have  a  gage 
cock  and  union  so  that  it  may  be  easily  removed.  The  operat- 
ing portion  of  the  gage  consists  of  a  flattened  copper  tube  bent 
in  a  circle  and  closed  at  the  end.  One  end  is  fixed,  or,  as  shown 
in  Fig.  29,  there  are  two  such  tubes.  When  fluid  pressure  is 


106 


HEAT  ENGINES 


applied  to  the  inside  of  the  tube,  its  cross-section  tends  to  assume 
a  circular  form  and  the  tube  tends  to  straighten.  The  greater 
the  pressure  the  more  the  straightening  of  the  tube.  By  proper 
mechanism  this  change  of  form  due  to  pressure  is  registered  on  a 
dial,  which  when  properly  calibrated  shows  the  pressure  in  the 
boiler. 

Fig.  30  shows  the  elevation  and  cross-section  of  a  water 
column  with  its  gage  glass.  The  section  shows  the  float  so 
arranged  that  it  will  blow  a  whistle  when  the  water  in  the 
boiler  is  too  high  or  too  low.  This  is  called  a  "high  and  low 
water  alarm." 


Elevation  Cross-section. 

FIG.  30. — Water  column. 

The  water  gage  and  water  column  to  which  it  is  attached 
are  important  accessories  in  boiler  operation.  The  length  of 
the  water  gage  on  the  boiler  should  be  such  as  to  cover  the  ordi- 
nary fluctuations  of  water  in  the  boiler.  It  should  always  be 
attached  to  a  water  column.  On  this  water  column  are  placed 
tri-cocks,  or  gage  cocks,  used  as  a  check  upon  the  water  column, 
as  the  water  column  is  sometimes  clogged  with  dirt.  The 
lowest  point  in  the  gage  glass  should  be  set  about  3  in.  above 


BOILERS 


107 


the  highest  point  of  the  tubes  in  tubular  boilers.  The  position 
of  the  gage  glass  in  water-tube  boilers  is  usually  determined 
by  the  manufacturers.  The  top  of  the  water  column  should 
be  attached  to  the  steam  space  so  that  it  will  get  dry  steam, 
and  the  bottom  of  the  water  column  to  the  water  space  at  a 
point  in  the  boiler.  There  should  be  blow-off  valves  on  both 
the  water  column  and  the  water  gage.  The  water  columns 
and  gage  cocks  should  be  blown  off  frequently.  Fig.  30  shows 
the  ordinary  arrangement  of  water  column, 
water  gage,  and  tri-cocks. 

Safety  valves  are  constructed  in  a 
great  many  different  forms,  but  in  general 
they  consist  of  a  valve  opening  outward 
and  held  in  place  by  a  spring,  and  in  the 
old  forms  by  an  arm  and  weight.  Fig. 
31  shows  the  construction  of  the  ordi- 
nary safety  valve.  The  size  of  the  safety 
valve  is  usually  determined  by  the  grate 
surface  and  the  steam  pressure  carried. 
The  following  rule  may  be  used : 

Let  G  =  the  grate  surface  in  square  feet; 
P  =  the    pressure   in   pounds    per 

square  inch  gage; 

A  =  the  total  area  of  safety  valve,  or  valves,  in  square 
inches. 

22. 5G 


FIG.  31. — Safety  valve. 


Then, 


AJL       —~       T»         * 


P  +  8.62 


Some  authorities  allow  in  spring-loaded  safety  valves  1 
sq.  in.  of  safety  valve  for  every  3  sq.  ft.  of  grate  surface. 
Formerly  the  lever  safety  valve  was  the  type  most  used,  but 
it  was  easily  tampered  with.  At  the  present  time  the  pop 
safety  valve  is  almost  universally  used.  Safety  valves  are 
adjusted  so  as  to  blow  at  one  pressure,  and  seat  at  a  pressure 
usually  2  Ibs.  less  than  that  at  which  they  open.  The  safety 
valve  on  the  boiler  should  be  tried  once  a  day  at  least,  to 
see  if  it  is  in  working  condition. 

In  an  article  presented  to  the  A.S.M.E.,  the  following  expres- 
sions have  been  developed  for  determining  the  size  of  safety 
valves  to  be  used  on  boilers: 


108  HEAT  ENGINES 

For  45°  valve  seats 

Z)  =  -0095ifp'- 

For  locomotives 


For  fire-tube  and  water-tube  stationary  boilers 

H 


J  LXP' 


For  marine  boilers 


E  =  Pounds  of  steam  discharged,  or  boiler-evaporation,  per 

hour. 

L  =  Vertical  lift  of  the  valve  in  inches. 
P  =  Steam  pressure  (absolute)  in  pounds  per  square  inch. 
D  =  Nominal  diameter  of  valve  (inlet)  in  inches. 
H  =  Total  boiler  heating  surface  in  square  feet. 

The  average  lift  for  a  safety  valve  is  about  .  1  of  an  inch. 
More  exact  results  may  be  obtained  by  reference  to  a  paper 
on  this  subject  by  P.  G.  Darling  in  the  A.S.M.E.  Proceedings 
for  1909. 

The  feed  pipe  to  the  boiler  is  always  provided  with  a  valve 
and  check  valve.  In  case  of  accident  to  the  feed  valve  the 
check  valve  will  close  and  prevent  the  water  from  leaving  the 
boiler. 

It  sometimes  happens  that  a  boiler  shell  may  become  over- 
heated and  a  boiler  explosion  results  from  this  accident.  Such 
accidents  are  avoided  by  having  screwed  into  the  boiler  a  plug 
consisting  of  a  brass  bushing  filled  with  a  metal,  which  melts 
before  any  damage  can  be  done  to  the  boiler.  These  plugs 
are  called  fusible  plugs  and  are  often  used. 


CHAPTER  VII 
BOILER  AUXILIARIES 

68.  Mechanical  Stokers. — In  firing  a  boiler  the  best  results 
are  obtained  by  firing  the  coal  in  small  quantities,  or  by  pro- 
gressive burning  of  the  coal.  With  hand  firing  these  results 
are  difficult  to  accomplish.  Most  firemen  prefer  to  shovel  the 


Transverse  Section 

FIG.  32. — Murphy  stoker — cross-section. 

coal  into  the  furnace  in  relatively  large  amounts  and  then  rest. 
It  is  difficult  to  get  them  to  give  the  proper  attention  to  the 
handling  of  their  fires.  With  mechanical  stokers  it  is  possible 
to  introduce  small  quantities  of  coal  frequently,  or  so  arrange 
the  stoker  that  there  may  be  progressive  burning  of  the  coal. 

The  first  stoker  was  introduced  into  England  by  Brunton 
in  1822,  and  at  nearly  the  same  time  by  Stanley.     These  were 

109 


110 


HEAT  ENGINES 


both  of  the  sprinkling  type.     The  first  chain  grate  was  brought 
out  by  John  Juckes.     The  first  American  stoker  was  invented 


FIG.  33. — Murphy  stoker — view  from  rear  of  furnace. 

by  Thomas  Murphy  of  Detroit,  Mich,  in  1878,  and  it  was  prod- 
bly  the  first  to  have  a  sloping  grate. 

Stokers   may   be   divided   into  .three   principal    classes:   the 
inclined  grate,  the  chain  grate,  and  the  under-feed. 


FIG.  34. — Elevation  of  Detroit  stoker. 

69.  Inclined  Grates. — The  Murphy  stoker,  shown  in  Figs.  32 
and  33,  is  an  example  of  the  inclined  grate  known  as  the  side 
overfeed  or  opposed  type. 


BOILER  AUXILIARIES 


111 


"At  either  side  of  the  furnace,  extending  from  front  to  rear,  is 
a  coal  magazine  into  which  the  coal  may  be  introduced  either 
mechanically  from  conveyors,  or  by  hand.  At  the  bottom  of  this 
magazine  is  the  coking  plate,  against  which  the  inclined  grates 
rest  at  their  upper  ends.  The  stoker  boxes,  operated  by  seg- 
ment gear  shaft  and  racks,  push  the  coal  out  over  the  coking 
plate  and  on  to  the  grates. 

The  grates  are  made  in  pairs — one  fixed,  the  other  movable. 
The  movable  grates,  pinioned  at  their  upper  ends,  are  moved  by 
a  rocker  bar  at  their  lower  ends,  alternately  above  and  below 
the  surface  of  the  stationary  grates.  The  stationary  grates 
rest  upon  the  grate  bearer,  which  also  contains  the  clinker  or 
ash  grinder.  This  grate  bearer  is  cast  hollow  and  receives  the 


FIG.  35. — Detroit  stoker — view  from  rear  of  furnace. 

exhaust  steam  from  the  stoker  engine.  This  steam  escapes 
through  small  openings  at  regular  intervals  on  either  side  of  the 
clinker  grinder  and  lower  ends  of  the  grates,  to  soften  the  clinker 
and  so  assist  the  cleaning  process." 

Figs.  34  and  35  show  the  elevation  and  rear  view  of  a  Detroit 
stoker.  This  is  very  similar  to  the  Murphy.  In  small  plants 
a  worm  conveyor  type  of  feed  is  used  (see  Fig.  35).  These 
conveyors  revolve  through  the  hoppers  and  feed  the  coal  in 
through  the  magazines.  This  is  a  convenient  arrangement  where 
the  coal  is  shoveled  into  the  hoppers  from  the  floor  by  hand. 

Both  the  Murphy  and  Detroit  stokers  are  adaptable  to  all 


112  HEAT  ENGINES 

grades  of  bituminous  coal,  but  are  not  suited  to  the  use  of  lignite 
or  anthracite. 

The  Dutch  oven,  or  extension  settings,  are  generally  re- 
garded as  the  most  effective  both  as  to  efficiency  and  as  to  the 
elimination  of  smoke. 

Still  another  form  of  inclined  grate  stoker  is  the  Roney, 
shown  in  Fig.  36,  in  whichTh!Tc7)aT7s~T^ 

of  the  boiler,  and  is  pushed  upon  an  inclined  grate  from  the 
front.     The  feeding  mechanism  is  shown  in  Fig.  37.     To  the 


FIG.  36. — Roney  stoker — side  view. 

main  operating  shaft  which  runs  horizontally  just  beneath  the 
hopper,  is  keyed  an  eccentric  which  gives  a  pendulum  motion  to 
the  agitator.  This  agitator  is  connected  with  the  grates,  giving 
to  them  a  rocking  movement.  It  also,  through  the  rock-shaft, 
imparts  a  reciprocating  motion  to  the  pusher,  the  length  of  this 
motion  being  regulated  by  a  hand- wheel.  As  the  pusher  recedes, 
the  fuel  in  the  hopper  settles  down  in  front  of  it,  and  as  it  ad- 
vances the  fuel  is  pushed  into  the  furnace.  As  the  main  operating 
shaft  runs  at  constant  speed,  the  quantity  of  fuel  fed  is  propor- 
tional to  the  travel  of  the  pusher. 

The  reciprocating  motion  of  the  rocker  bar  is  imparted  through 


BOILER  AUXILIARIES 


113 


a  connecting  rod,  the  free  end  of  which  works  freely  through  a 
sliding  bearing  in  the  lower  end  of  the  agitator — an  adjusting 
nut  on  the  connecting  rod  makes  it  possible  to  regulate  the  re- 
ciprocating motion  of  the  rocker  bar,  and  consequently  the 
amplitude  of  the  rocking  motion  of  the  grate  bars. 

The  inclined  grate  stoker  has  given  excellent  satisfaction, 
particularly  in  using  diversified  coals.  In  conditions  of  excessive 
loads  this  is  probably  not  so  smokeless  as  some  other  forms 
of  stoker,  but  when  carefully  operated  is  one  of  the  most  satis- 
factory forms. 


hopper  Plate 


Feed  Wheel-., 
Agitator-Sector 


Sheath-Nut 

Hock- Shaft 

Agitator.. 
Lock- Nut  - 


Main  Operating  Shaft--- 
Connecting  Rod- 


Rocker  Bar- 


FIG.  37. — Roney  stoker — feed  mechanism. 

70.  Chain  Grates. — Fig.  38  shows  the  elevation  of  a  chain 
grate.  The  coal  is  fed  into  a  hopper,  the  bottom  of  which  is 
open  to  the  chain  grate,  composed  of  a  series  of  flexible  links 
rotating  upon  two  cylinders,  one  at  each  end  of  the  grate. 
The  grate  is  driven  by  a  small  engine  the  speed  of  which  can 
be  adjusted  to  the  particular  form  of  fuel  burned  and  the  load 
on  the  boiler.  This  speed  should  be  regulated  so  that  the  fuel 
is  completely  burned  just  as  it  reaches  the  back  of  the  grate.  If 


114 


HEAT  ENGINES 


the  speed  is  too  fast,  unburned  coal  will  be  carried  over  the  back 
of  the  grate  into  the  ash  pit.     If  it  is  too  slow,  there  will  be  holes 


FIG.  38. — Green  chain  grate. 


FIG.  39. — Green  chain  grate  stoker  applied  to  Stirling  boiler. 

in  the  fire  toward  the  rear,  allowing  an  excess  of  air  to  pass 
through  the  grate.  The  coal  drops  from  the  hopper  upon  this 
slowly  moving  grate,  the  thickness  of  the  bed  of  coal  being  ad- 


BOILER  AUXILIARIES  115 

justed  by  an  apron  at  the  front  of  the  boiler.  This  form  of  grate 
gives  excellent  satisfaction  with  non-coking  coals  and  uniform 
loads  on  the  boiler,  and  will  be  almost  smokeless  under  proper 
conditions  of  operation.  It  is  not  adapted  to  the  use  of  semi- 
bituminous,  or  anthracite,  coal.  The  greatest  difficulty  is  im- 
proper installation,  which  will  permit  of  the  passing  of  an  excess 
of  air  through  the  grate.  This,  however,  may  be  avoided  by 
careful  setting.  In  installing  these  grates,  provision  should  be 
made  for  the  easy  removal  of  the  ashes.  Fig.  39  shows  the 


FIG.  40. — Jones  under-feed  stoker. 

cross-section  of  a  chain  grate  installed  under  a  Stirling  boiler. 
It  also  shows  the  ash  pit  and  sub-basement  for  easy  removal 
of  the  ashes.  .This  is  a  desirable  arrangement  with  most  forms 
of  stokers. 

71.  Under-feed  Stokers. — One  of  the  commonest  forms  of 
under-feed  stokers  is  the  Jones,  shown  in  Fig.  40  applied  to 
a  boiler  plant,  and  in  Fig.  41  in  cross-section.  In  this  form, 
coal  is  dropped  down  from  hoppers  in  front  of  a  piston  at  regular 
intervals  depending  upon  the  load.  This  piston  moves  forward 


116 


HEAT  ENGINES 


and  pushes  the  coal  in  under  the  burning  fuel.     In  this  way 
coal  is  always  introduced  under  the  fire,  and  all  the  gases  are 


FIG.  41. — Section  of  Jones  under-feed  stoker. 

passed  through  the  incandescent  fuel.     This  is  the  most  smokeless 
form  of  stoker. 

Fig.  42  shows  the  American  type  of  stoker,  which  is  similar 


FIG.  42. — American  under-feed  stoker. 

in  operation,  but  in  this  stoker  the  piston  of  the  Jones  is  replaced 
by  a  worm  which  continuously  feeds  the  coal  underneath  the  fire. 


BOILER  AUXILIARIES 


117 


The  under-feed  form  of  stoker  produces  a  very  intense  heat 
directly  above  the  fire.  The  ash  accumulates  on  the  top  of  the 
fire  and  falls  over  to  the  sides  of  the  furnace,  from  where  it  is 
taken  out.  Owing  to  the  ashes  being  raised  to  a  high  tempera- 
ture, coal  containing  ash  which  is  high  in  sulphur  and  iron  should 
not  be  used  in  a  stoker  of  this  type,  as  it  will  produce  very  large, 
hard  clinkers. 

In  under-feed  stokers,  the  resistance  of  the  fuel  bed  to  the 
passage  of  air  is  so  great  that  it  is  necessary  to  use  a  blower 


FIG.  43. — Taylor  stoker — cross-section. 

to  force  the  air  through  the  fuel  bed.  This  blower  is  usually 
driven  by  a  steam  engine,  and  the  excessive  amount  of  steam  used 
is  one  of  the  objections  offered  to  the  use  of  these  stokers. 

The  Taylor,  Figs.  43  and  44,  is  an  inclined  under-feed 
stoker. 

"Coal  from  the  hopper  is  fed  into  the  retort,  from  which  two 
cylindrical  rams,  assisted  by  gravity,  introduce  it  into  the  fur- 
nace at  an  angle  to  the  fire  surface.  Movement  of  the  upper 


118 


HEAT  ENGINES 


ram  pushes  the  green  coal  outward  and  upward,  properly  dis- 
tributing it  in  the  coking  zone.  The  action  of  the  lower  ram  is 
similar,  but  instead  of  bringing  in  fresh  coal  it  pushes  the  fuel 
bed  and  refuse  toward  the  dump  plates  at  the  rear. 

The  retort  or  fuel  magazine  is  formed  by  two  tuyere  boxes. 
Air  for  combustion  enters  the  tuyere  boxes  from  the  wind  box, 
and  escaping  from  the  tuyere  openings  mingles  with  the  gases 
distilled  from  the  coal  and  with  the  coked  fuel  pushed  outward 
and  upward  by  the  rams.  Both  rams  are  actuated  by  connecting 
rods  and  links  from  a  crank  shaft  which  is  driven  from  the  speed 
shaft.  The  speed  shaft  in  turn  is  driven  by  the  fan  engine. 


FIG.  44. — Taylor  stoker — perspective,  showing  dumping  plates  down. 

"The  dump  plates,  which  are  combination  dump  plates  and 
fire  guards,  are  hung  on  the  rear  of  the  wind  box;  these  plates 
receive  the  burned  out  refuse  and  are  dumped  periodically,  as 
the  conditions  of  service  may  require.  The  dump  plates  are 
operated  from  the  front  of  the  stoker,  raised,  latched  in  position 
and  released  by  a  hand  lever." 

72.  Grate  Surface  in  Stokers. — The  grate  surface  in  a  stoker 
with  an  inclined  grate  is  taken  as  the  area  of  the  horizontal 
projection  of  the  grates,  and  is  termed  projected  area.  The 
ratio  of  projected  grate  area  in  the  stoker  to  the  heating  surface 
in  the  boiler  varies  from  1  to  55,  to  1  to  65. 


BOILER  AUXILIARIES 


119 


73.  Advantages  and  Disadvantages. — The  principal  advan- 
tages of  mechanical  stokers  are:  smokeless  operation  of  the 
furnace,  adaptability  to  the  burning  of  cheaper  grades  of  coal, 
uniformity  of  furnace  conditions  and  steam  pressure,  which  adds 
to  the  economy  of  the  plant,  and  in  larger  plants,  a  saving  in 
the  labor  charge  for  plant  operation.  Their  disadvantages  are: 
high  initial  cost,  large  repair  bills,  cost  of  operating  stoker 
mechanism,  which  in  most  stokers  is  from  J  to  3  per  cent,  of 
the  steam  generated,  and,  if  fan  blast  is  used,  from  3  to  5  per 
cent,  of  the  steam  generated. 


FIG.  45. — General  arrangement  of  a  modern  boiler  room. 

In  small  plants  where  coal-handling  machinery  is  not  pro- 
vided, stokers  will  not  reduce  the  labor  charge.  In  large  plants 
where  the  coal  is  delivered  mechanically  to  the  stoker  hoppers, 
stokers  will  materially  reduce  this  charge.  Fig.  45  shows  a  plant 
with  stokers  fed  from  overhead  hoppers.  In  such  plants  the 
ash  is  usually  removed  from  a  basement  under  the  boiler  room 
floor. 

74.  Boiler  Feed  Pumps. — The  feed  water  is  forced  into  a 
boiler  either  by  a  feed  pump  or  an  injector.  There  are  wot 
general  types  of  feed  pumps:  belted  feed  pumps  which  may 


120 


HEAT  ENGINES 


be  driven  from  the  machinery,  or  by  independent  motor;  and 
independent  pumps  driven  by  their  own  steam  cylinders. 

The  independent  feed  pump  is  most  commonly  used  as  it 
has  the  advantage  of  being  independent  of  the  operation  of  the 
main  engine,  and  in  addition  its  speed  can  be  adjusted  so  as  to 
give  uniform  feeding.  Its  principal  disadvantage  is  in  the  large 
steam  consumption  of  pumps  of  this  type.  Small  feed  pumps 
use  from  150  to  300  Ibs.  of  steam  per  indicated  horse-power  per 
hour;  large  steam  pumps,  from  80  to  150  Ibs.;  compound  con- 
densing feed  pumps  of  the  direct-acting  type,  from  60  to  75  Ibs. 
The  mechanical  efficiency  of  these  pumps  is  about  80  per  cent. 


FIG.  46. — Worthington  boiler  feed  pump. 

Fig.  46  shows  a  modern  form  of  feed  pump  having  four  single- 
acting  water  cylinders.  This  pump  has  -two  plungers  working 
in  these  cylinders.  The  plungers  are  in  the  center  of  the  pump 
and  have  the  packing  glands  outside  the  cylinder.  This  type 
of  pump  is  called  an  outside  center  packed  pump. 

The  belt-driven  pump  is  often  used  to  overcome  the  steam 
wasted  when  using  the  independent  direct-acting  pump.  These 
pumps  may  be  driven  from  the  shaft  of  the  main  engine  or  from 
the  line  shafting.  In  some  cases  they  are  driven  by  an  electric 
motor.  This  arrangement  has  its  disadvantages.  The  speed  of 
the  pump  being  constant,  it  is  necessary  to  regulate  the  amount 
of  water  pumped  by  a  by-pass  allowing  part  of  the  water  pumped 
to  go  back  from  the  pressure  to  the  suction  side  of  the  pump. 


BOILER  AUXILIARIES 


121 


If  the  feed  is  suddenly  shut  off  from  all  the  boilers,  provision 
must  be  made  for  the  discharge  from  the  pump  being  turned  back 
to  the  suction  automatically.  It  is  not  possible  to  use  a  belted 
feed  pump  except  when  the  engine  is  running,  and  there  must 
be  an  auxiliary  feeding  device  provided  that  can  be  operated 
when  the  main  engine  is  shut  down. 

In  very  large  plants  steam  turbine  driven,  turbine  pumps  are 
being  used.  These  pumps  being  of  the  centrifugal  type,  it  is 
not  necessary  to  change  the  speed  of  the  pump  for  changes  of 
load.  The  speed  of  a  turbine  pump  determines  the  pressure 


Overflow 

Water 
FIG.  47. — Steam  injector — cross  section. 

only  and  the  amount  of  water  pumped  depends  upon  the  demand. 
This  is,  therefore,  automatic  and  requires  very  little  attention. 

75.  Steam  Injectors. — Boilers  are  often  fed  by  an  injector, 
a  device  invented  by  M.  Giffard,  a  French  engineer. 

Fig.  47  shows  the  cross-section  of  an  injector,  the  operation 
of  which  is  as  follows:  The  handle,  137,  is  pulled  back  slightly, 
thus  raising  valve  130  from  its  seat  and  admitting  steam  through 
valve  126  to  the  lifter  nozzel  101.  "The  discharge  of  steam 
from  this  nozzle  into  the  lifter  combining  tube,  102,  entrains 
the  air  in  the  suction  pipe  finally  producing  sufficient  vacuum 
to  lift  the  water.  The  flow  of  water  passes  through  both  the 
intermediate  overflow,  121,  and  the  forcer  combining  tube, 
104,  and  out  of  the  final  overflow,  117.  A  further  movement 
of  the  lever  opens  the  forcer  steam  valve,  126,  and  admits 


122  HEAT  ENGINES 

steam  to  the  forcer  steam  nozzle,  103,  while  at  the  same  time 
the  final  overflow  valve  is  approaching  its  seat,  producing  a 
consevuent  increase  of  pressure  in  the  delivery  chamber.  This 
pressure  closes  the  intermediate  overflow  valve,  121,  and  opens 
the  intermediate  or  line  check  valve,  111,  and  when  the  final 
overflow  valve,  117,  is  brought  to  its  seat  the  injector  will  be 
in  full  operation.  The  intermediate  overflow  valve,  117, 
operates  automatically,  its  only  function  .being  to  give  direct 
relief  to  the  lifter  steam  nozzle  when  lifting  or  priming,  and 
comes  to  its  seat  when  the  forcer  steam  is  applied  and  is  held 
there  by  the  pressure  exerted  by  the  forcer." 

There  are  many  different  forms  of  injectors  made  for  different 
conditions.  The  injector,  however,  is  a  very  inefficient  pump 
for  general  pump  purposes.  It  is  installed,  however,  as  an  auxil- 
iary method  of  feeding  the  boiler  in  case  of  accident  to  the  regular 
feed  pump.  As  a  boiler  feeder  it  has  a  thermal  efficiency  of  al- 
most 100  per  cent,  since  all  the  heat  of  the  steam  used  by  the  in- 
jector, except  that  lost  by  radiation,  goes  into  the  feed  water. 

In  locomotives,  injectors  only  are  used  for  feeding  the  boiler, 
as  they  take  very  little  space  and  warm  the  feed  water.  Each 
locomotive  is  provided  with  two  injectors. 

76.  Pump  Connection. — When   a  pump  or   injector   is   han- 
dling cold  water,  the  lift  on  the  suction  side  of  it  should  not 
exceed  25  ft.     Most  engineers  try  to  install  pumping  apparatus 
with  a  head  on  the  suction  not  more  than  15  ft. 

When  hot  water  is  to  be  handled,  the  pump  should  be  below 
the  level  of  the  water  on  the  suction  side.  By  hot  water  is 
meant  water  exceeding  120°.  Injectors  are  seldom  used  to 
handle  hot  water  as  they  are  very  difficult  to  start  with  water 
exceeding  100°.  Where  pumps  are  installed  handling  hot  water 
from  a  feed-water  heater,  the  level  of  water  in  the  heater  should 
be  5  ft.  above  the  center  line  of  the  pump  cylinders  if  pos- 
sible. Hot  water  cannot  be  raised  by  a  pump,  as  the  lowering 
of  the  pressure  in  the  suction  pipe  lowers  the  temperature  of 
the  boiling  point  of  the  water  in  the  suction  pipe,  the  water 
in  the  suction  boils  and  all  the  pump  draws  from  the  suction 
is  steam. 

77.  Feed-water  Heaters. — It  is  very  important  that  a  boiler 
be  fed  with  warm  water,  usually  at  a  temperature  over  180°. 
This  saves  part  of  the  heat  necessary  to  make  steam,  and  in 
addition  prevents  strains  in  the  boiler  due  to  a  difference  in 


BOILER  AUXILIARIES  123 

temperature  of  different  parts  of  the  boiler  shell.  Feeding  a 
boiler  with  cold  water  often  causes  a  leak. 

In  all  modern  power  plants  some  means  is  provided  for 
heating  the  feed  water  before  entering  the  boilers.  This  is 
usually  accomplished  in  one  of  two  ways;  by  heating  the  water 
with  the  exhaust  steam  from  the  engine,  which  is  by  far  the 
commonest  method  used,  or  with  waste  gases  from  the  boilers. 
Devices  for  using  the  exhaust  steam  for  heating  the  water  are 
called  feed-water  heaters,  and  the  device  for  using  the  gases  from 
the  boiler  for  heating  the  feed  water  is  termed  an  economizer. 

The  principal  advantages  of  the  feed-water  heater  are  the 
saving  in  B.T.U.  due  to  the  increase  'in  the  temperature  of  the 
feed,  and  the  saving  in  wear  and  tear  on  the  boiler  due  to  in- 
troducing hot  instead  of  cold  water,  thereby  reducing  the 
strain  on  the  boiler.  A  heater  which  increases  the  temperature 
of  the  feed  water  from  70°  to  200°  will  save  about  12  per  cent, 
of  the  fuel,  and  the  installation  of  a  heater  will  usually  pay  for 
itself  in  a  few  months. 

78.  Types  of  Feed-water  Heaters. — There  are  two  general 
types  of  heaters:  the  open  and  the  dosed.  The  open  feed- 
water  heater,  Fig.  48,  consists  of  a  cast-  or  wrought-iron  shell 
into  which  the  exhaust  steam  is  led.  The  cold  water  is  admitted 
at  the  top  of  the  heater,  and  is  allowed  to  pass  through  the 
exhaust  steam  in  streams  or  sheets  of  water.  In  this  type  of 
heater  the  feed  water  and  exhaust  steam  come  into  direct  con- 
tact with  each  other.  The  water  usually  passes  over  pans,  or 
trays,  upon  which  any  scale-producing  matter  can  be  deposited. 
When  it  is  desired  to  clean  the  heater,  it  is  only  necessary  to 
take  out  these  pans  and  clean  them.  Before  entering  the  heater 
the  exhaust  steam  should  be  passed  through  an  oil  separator. 
The  hot  feed  water  is  usually  passed  through  some  form  of 
filter  before  going  to  the  feed  pumps.  The  feed-water  heater 
should  be  located  at  a  sufficient  height  above  the  feed  pump  so 
that  the  water  will  enter  at  a  pressure.  This  distance  should 
be  5  ft.  or  more.  The  heater  may  also  be  used  as  a  receptacle 
for  the  hot  water  which  is  drained  from  the  steam  mains,  and 
for  other  hot  condensed  steam  which  does  not  contain  oil.  A 
uniform  water  level  is  maintained  in  the  heater  by  a  float  valve 
which  automatically  allows  water  to  enter  the  heater  when  the 
level  gets  below  a  certain  point. 

The  closed  heater  shown  in  Fig.  49  consists  of  a  cylindrical 


124 


HEAT  ENGINES 


shell  of  cast  iron,  or  steel,  containing  tubes  extending  from  the 
header  at  one  end  of  the  heater  to  the  header  at  the  other  end, 


FIG.  48. — Open  feed-water  heater. 

or  tubes  in  the  form  of  coils  of  pipe.  The  exhaust  steam  is 
admitted  on  one  side  of  the  tubes  and  the  feed  water  on  the 
other.  In  a  closed  heater  the  feed  water  and  the  steam  used 


Blow 


Feed 


Steam  Outlet 

or  Inlet 


FIG.  49. — Closed  feed -water  heater. 

do  not  come  in  contact  with  each  other.     The  closed  heaters 
are  usually  used  where  it  is  desired  to  pass  the  water  through 


BOILER  AUXILIARIES 


125 


the  heaters  under  pressure.  They  are  more  expensive  than  the 
open  heaters  and  are  more  difficult  to  clean.  Where  possible 
it  is  better  to  use  an  open  heater. 

79.  Installation  of  Heaters. — Open  heaters  are  placed  on  the 
suction  side  of  the  feed  pump,   and  the  feed  water  must  be 
brought  to  the  heater.     The  level  of  the  water  in  an  open  heater 
should  be  at  least  5  ft.  above-  the  center  of  the  feed-pump  cylin- 
der as  a  feed  pump  cannot  lift  hot  water.     Injectors  are  never 
used  with  an  open  heater  as  they  cannot  be  used  with  hot  water. 

Closed  heaters  are  placed  on  the  discharge  side  of  the  pump 
and  the  feed  pump  may  lift  its  supply  directly  from  the  source 
of  water.  An  injector  may  be  used  with  a  closed  heater. 

Heaters  cost  from  $2  to  $4  per  boiler  horse-power  served  by 
them. 

80.  Economizers. — Any  device  which  heats  the  feed  water  by 
means  of  the  heat  in  the  gases  which  leave  the  boiler  is  termed 


End  elevation.  Side  elevation. 

FIG.  50. — Economizer. 

an  economizer.  Fig.  50  shows  the  elevations  of  an  economizer. 
The  cold  water  is  pumped  into  the  lower  pipe  header,  and  after 
being  heated,  passes  out  from  the  upper  header  to  the  boiler. 
The  flue  gases  from  the  boiler  pass  around  the  pipes  and  headers 
containing  the  feed  water.  The  tubes  as  shown  in  the  cut  are 
provided  with  scrapers  operated  from  time  to  time  to  remove 
the  soot  from  the  pipes.  The  general  arrangement  of  an  econ- 
omizer is  shown  in  Fig.  51.  An  economizer  is  always  provided 


126 


HEAT  ENGINES 


with  a  duct,  or  by-pass,  passing  around  it,  so  that  it  can  be 
cleaned  without  shutting  down  the  plant.  The  economizer  is 
placed  in  a  brick  or  sheet  metal  flue  which  carries  the  gases 
from  the  boiler  to  the  chimney.  Economizers  are  installed  so 
as  to  make  use  of  the  heat  in  the  gases  leaving  a  boiler  and  thus 
reduce  the  waste  in  heat  going  up  the  stack.  Economizers  may 
be  installed  also  to  increase  the  capacity  of  a  boiler  plant  which 
is  too  small  for  its  services.  They  deliver  the  water  to  the 


FIG.  51. — Economizer,  showing  location  in  breeching. 

boiler  at  a  high  temperature,  reducing  the  strain  and  the  leakage 
caused  by  the  admission  of  cold  water.  Their  particular  dis- 
advantage is  in  reducing  the  strength  of  the  draft  owing  to  the 
fact  that  the  economizer  causes  additional  friction.  Econo- 
mizers should  never  be  used  except  with  chimneys  having  a 
strong  draft  or  with  mechanical  draft. 

The  first  cost  of  the  economizer  is  very  high,  varying  from 
$5  to  $6  per  horse-power  for  plants  of  1000  horse-power  or  over. 
A  number  of  tests  have  been  made  of  the  economizer  where  a 


BOILER  A  UXILI ARIES  127 

net  saving  of  10  per  cent,  was  shown,  allowing  for  cost  of  econo- 
mizer, cost  of  operation,  interest,  depreciation,  and  repairs. 

From  4  to  5  sq.  ft.  of  economizer  surface  should  be  allowed 
per  boiler  horse-power. 

81.  Superheaters. — In  the  past  few  years  the  use  of  super- 
heated steam  with  both  reciprocating  engines  and  turbines  has 
become  very  general.  The  benefits  derived  are  many.  The 
steam  remains  in  a  dry  condition  until  all  the  superheat  is  lost. 
The  heat  lost  by  the  steam  while  passing  through  the  piping 
from  the  superheater  to  the  place  where  it  is  to  be  used,  does  not 


B 


FIG.  52. — Superheating  coil  in  Babcock  and  Wilcox  boiler. 

cause  condensation  as  it  is  simply  superheat  which  is  given  up. 
The  initial  condensation  loss  in  reciprocating  engines  is  greatly 
reduced,  or  entirely  eliminated,  depending  upon  the  amount 
of  superheat  in  the  steam.  In  turbines  the  absence  of  moisture 
is  particularly  desirable,  as  the  water  coming  in  contact  with 
the  blading  at  a  high  velocity  has  an  eroding  effect,  thus  in- 
creasing the  clearance  between  the  blades  and  the  casing  and 
consequently  increasing  the  steam  consumption. 

Recent  experiments  have  shown  that  when  steam  is  super- 
heated from  0°  to  100°  F.  there  is  a  saving  of  1  per  cent,  in 


128  HEAT  ENGINES 

steam  consumption  for  every  10  degrees  of  superheat,  and 
when  superheated  from  100°  to  200°  F.  there  is  a  saving  of 
1  per  cent,  for  every  12  degrees  of  superheat.  These  results  are 
based  on  a  comparison  between  superheated  and  dry  saturated 
steam.  If  the  steam  is  wet.  the  saving  will,  of  course,  be  much 
larger. 

The  degree  of  superheat  to  be  used  will  depend  largely  upon 
the  conditions.  In  the  majority  of  cases  it  has  been  found  that 
the  highest  commercial  efficiency  is  secured  by  the  use  of  from 
125°  to  150°  of  superheat  in  turbine  plants  and  slightly  less  in 
the  case  of  reciprocating  engine  plants. 

A  superheating  coil  placed  in  a  Babcock  &  Wilcox  boiler  is 
shown  m  Figure  52. 

It  has  been  frequently  stated  that  cast-iron  fittings  and  valves 
should  not  be  used  with  superheated  steam,  as  the  iron  deterio- 
rated at  the  high  temperatures.  Recent  developments  have 
shown  that  the  trouble  has  been  caused  by  fluctuating  rather  than 
high  temperatures. 

In  the  transactions  of  the  A.S.M.E.,  Vol.  31,  page  1037, 
Professor  Hollis  states: 

"When  the  temperature  is  constant,  even  though  as  high  as  600°  or 
700°  F.,  the  change  in  cast  iron  is  not  serious  enough  to  prohibit  us  from 
its  use,  but  where  the  temperature  varies  considerably,  the  metal  is 
certain  to  develop  cracks  and  distortion  that  render  it  unsuitable  for 
steam  pipes  and  other  parts  under  steam  pressure. 

"The  use  of  cast-iron  fittings  for  superheated  steam  is  inadvisable 
where  the  temperature  is  likely  to  fluctuate,  but  it  can  be  safely  used 
where  the  temperature  is  to  be  constant." 

82.  Chimneys. — The  chimney  is  a  very  important  part  of  a 
steam-power  plant,   and   the   operation  of  the  plant   depends 
upon  the  draft  and  capacity  of  the  chimney. 

83.  Draft. — The  draft  in  a  chimney  is  produced  by  the  dif- 
ference in  weight  between  the  column  of  hot  gases  inside  the 
chimney  and  a  column  of  gases  of  the  same  dimensions  outside 
the  chimney.     The  hot  gases,   being  light,   are  forced  up  the 
chimney  by  the  cold  gases  coming  through  the  grates. 

The  height  of  the  chimney  then  determines  the  intensity  of 
the  draft.  The  draft  is  always  measured  in  inches  of  water, 
and  for  a  given  height  of  stack  may  be  determined  as  follows: 


BOILER  AUXILIARIES  129 

Let   H  =  the  height  of  the  chimney. 

T°  =  the  absolute  temperature  of  the  gases  outside  the 

chimney. 
T'  =  the  absolute  temperature  of  the  gases  inside  the 

chimney. 
w°  =  the  weight  of  a  cubic  foot  of  air  at  a  temperature 

/T7O 

w'  '=  the  weight  of  a  cubic  foot  air  at  a  temperature 

T'. 

Then  assuming  the  chimney  to  have  an  area  of  1  sq.  ft.,  the 
weight  of  the  hot  gases  equals 

Hw'  =  Hw°  ~r  (1) 

The  weight  of  the  cold  gases  equals 

Hw°  =  Hwe  ~  (2) 

Hence  the  force  of  the  draft, 

rpo 

F'  =  Hw°  -  Hw'  =  Hw°  -  Hw°  7^ 
Therefore 

F'  =  Hw°(l  -  J°)-  (3) 

This  is  in  pounds  per  square  foot.  To  reduce  to  inches  of 
water  this  must  be  multiplied  by  .192.  Hence  the  force  of  the 
draft  in  inches  of  water, 


F  =  .W2Hw°l  -  ~-  (4) 

The  intensity  of  the  draft  as  shown  in  equation  (4)  is  determined 
by  the  height  of  the  chimney  and  the  temperature  inside  and 
outside  the  chimney. 

84.  Chimney  Capacity.  —  -The  capacity  of  a  chimney  is  the 
quantity  of  gases  that  it  will  pass  per  hour,  and  upon  the  capacity 
of  a  chimney  depends  the  number  of  pounds  of  coal  that  the  plant 
will  burn.  The  theoretical  quantity  of  coal  that  a  chimney 
will  burn  may  be  found  as  follows: 

Let  h  =  the  head  producing  velocity.  Then  the  weight  of 
the  gases  producing  the  head  equals  hw',  and 

hw'  =  Hw°  -  Hw'  =  Hw'  ~  -  Hw'.  (5) 


130  HEAT  ENGINES 

Therefore 

IT'          \ 

h  =  H  (^0  -  1.)  (6) 

Let  u°  =  the  velocity  of  the  entering  gases  and  u'  =  the 
velocity  of  the  leaving  gases  in  feet  per  second.  Then  the  veloc- 
ity of  the  leaving  gases 


(7) 

Let  W°  =  the  total  weight  of  the  gases  passing  up  the  chimney 
per  second,  then 

W°  =  w°u°  =  w'u'  -- 


or 


For  an  outside  temperature  of  70°  F.,  w°  =  .075  and  T°  = 
530.  Assume  the  temperature  in  the  chimney  to  be  500°  F. 
Then  T'  equals  960°. 

Substituting  these  values  in  equation  (8), 


=  .602  #  X  .247.  (9) 

If  A  =  area  of  the  chimney  in  square  feet,  then 

W°  =  .30  A  V#  in  pounds  per  second,  (10) 

or  in  pounds  per  hour 

TF°!  =  3600  X  .3  A  Vff.  (11) 

This  assumes  the  efficiency  of  a  chimney  to  be  1,  but  experi- 
ence shows  the  average  efficiency  of  a  chimney  to  be  about  35 
per  cent.,  so  that  the  actual  weight  of  air  passed  per  hour  is 

W°a  =  3600  X  .35  X  .3A  V#  =  3784  V#.  (12) 

Each  pound  of  coal  requires  24  Ibs.  of  air  to  burn  it,  and  as 
each  boiler  horse-power  requires  about  5  Ibs.  of  coal,  the  boiler 
horse-power  of  a  chimney  is 


B.H.P.  =  ^  VH  =  3.15A  V#-  (13) 


BOILER  AUXILIARIES  131 

Various  authors  give  values  of  the  constant  in  this  expression 
varying  from  3.5  to  3.0. 

85.  Height  of  a  Chimney. — The  height  of  a  chimney  is  always 
measured  from  the  level  of  the  grate  and,  in  any  given  case, 
depends  upon  the  kind  of  fuel  that  is  to  be  burned  under  the  boiler. 
The  following  table  gives  the  minimum  height  of  chimney  for 
various  kinds  of  fuels : 

TABLE  XVIII.  CHIMNEY  HEIGHTS 

For  straw  or  wood 35  feet. 

bituminous  lump,  free  burning  100  " 

ordinary  slack 100 

ordinary  bituminous  coal 115  *' 

small  slack  or  anthracite 125  " 

anthracite  pea  coal 150  " 

The  height  of  the  chimney  should  not  be  too  short  for  its 
diameter.  A  very  large  diameter  of  chimney  in  proportion  to 
the  height  may  show  reduced  capacity.  As  an  example,  a  chim- 
ney 100  ft.  high  should  not  exceed  6.5  ft.  in  diameter.  In  general 
the  inside  diameter  of  a  chimney  should  not  exceed  8  per  cent,  of 
its  height. 

86.  Materials  Used. — Brick  or  hollow  tile  is  more  extensively 
used  in  building  chimneys  than  any  other  material  where  per- 
manent chimneys  are  desired.     The  life  of  a  brick  chimney  is 
probably  forty  or  fifty  years.     These  materials   are  used  in 
plants  where  few  changes  are  expected. 

In  most  plants  the  station  is  not  expected  to  remain  without 
extensive  changes  more  than  twenty  or  twenty-five  years,  and 
the  expense  of  a  brick  chimney  is  not  warranted.  Many  of 
the  recent  power  houses  are  using  self-sustaining  steel  chimneys. 

For  temporary  use  the  unlined  sheet  steel  chimney  is  very 
commonly  used.  It  is  necessary  to  brace  these  chimneys  with 
steel  guy  wires.  The  life  of  these  chimneys  is  short,  at  the  best 
not  more  than  ten  years,  and  where  the  coal  contains  much 
sulphur  not  more  than  five  years. 

87.  Brick  Chimneys. — Brick  chimneys,  as  shown  in  Fig.  53, 
are  built  in  two  parts,  an  outer  shell  and  an  inner  shell,  usually 
lined  with  fire-brick  which  forms  a  flue  for  the  burned  gases. 
There  should  be  an  air  space  between  the  outer  and  the  inner 
shells  so  that  the  inner  shell  is  free  to  expand.     Brick  chimneys 
are  expensive  to  erect,  but  very  permanent  in  character.     Care 
should  be  taken  in  investigating  the  ground  which  is  to  support 


132 


HEAT  ENGINES 


a  chimney,  as  unequal  or  excessive  settlement 
may  endanger  the  chimney. 

The  radial  brick  chimney  is  constructed 
of  hollow  tile  and  has  no  lining.  These  chim- 
neys are  much  lighter  than  the  solid  brick 
chimney.  They  are  much  less  expensive  than 
the  brick  and  cost  but  little  more  than  a  self- 
sustaining  steel  chimney. 

88.  Steel  Chimneys.— Steel  chimneys  of  the 
self-sustaining  type  are  built  of  boiler  plates 
riveted  together.     They  are  supported  on  am- 
ple foundations  to  which  they  are  bolted  by 
very  heavy  anchor  bolts.      The  pressure  of 
the  wind  -  against  the  chimney  is  carried  to 
the  foundation  by  these  bolts,  and  the  foun- 
dation must  be  of  sufficient  size  and  weight 
to   prevent    overturning.     Chimneys  of   this 
type  are  lined  with  fire-brick  usually  for  their 
full  length. 

89.  Mechanical  Draft. — In  some  cases  con- 
ditions will  not  permit  of  the  construction  of 
a  tall  chimney,  and  in  other  cases  the  draft 
required  is  more  than  the  ordinary  chimney 
will  give.     It  is  then  necessary  to  resort  to 
some  form  of  forced  or  mechanical  draft. 

Mechanical  draft  is  entirely  independent  of 
the  temperature  inside  or  outside  of  the  chim- 
ney. Where  economizers  are  used,  the  tem- 
perature in  the  chimney  may  be  so  low  and  the 
resistance  of  the  economizer  such  as  to  require 
mechanical  draft.  • 

90.  Systems  of  Mechanical  Draft. — There 
are  three  systems  that  may  be  used  to  produce 
mechanical  draft. 

(1)  A  steam  jet  may  be  used  to  force  air 
into  the  ash  pit. 

(2)  A  fan  may  be  used  to  force  air  into  the 
FIG.  53.— Brick           ^     ., 

chimney.  ash  P1^ 

Both  of  the  above  systems  require  a  closed 

ash  pit  and  are  termed  forced  draft,  as  the  air  is  forced  through 
the  fire. 


CC 


BOILER  AUXILIARIES  133 

(3)  The  third  system,  or  induced  draft,  is  more  commonly 
used.  With  this  system  a  fan  is  placed  in  the  smoke  connection 
to  the  chimney,  or,  as  in  the  case  of  locomotives,  the  cylinders 
exhaust  directly  into  the  stack,  and  air  is  drawn  through  the  fire. 
The  action  in  this  case  is  analogous  to  the  action  of  the  chimney. 

Under  ordinary  conditions  the  rate  of  combustion  may  be  taken 
as  from  15  to  30  Ibs.  of  coal  per  square  foot  of  grate  surface  per 
hour  with  mechanical  draft.  With  mechanical  draft  the  air 
required  to  burn  a  pound  of  coal  may  be  reduced  to  18  Ibs. 
With  induced  draft  the  pressure  of  the  draft  usually  varies  from 
1.5  to  2  in.  of  water.  The  operation  of  an  induced  draft 
plant  may  be  made  partially  automatic.  This  is  done  by  driv- 
ing the  fan  with  an  engine  and  having  the  speed  of  the  engine 
controlled  by  the  steam  pressure  in  the  boilers. 

PROBLEMS 

1.  Calculate  the  factor  of  evaporation  for  a  gage  pressure  of  75  Ibs.  and 
an  initial  temperature  of  the  feed  water  of  135°. 

2.  A  boiler  evaporates  500  Ibs.  of  water  per  hour  from  a  feed  temperature- 
of  145°  into  steam  at  80  Ibs.  pressure.     What  is  the  equivalent  water  evapo- 
rated per  hour  from  and  at  212°? 

V3.  A  boiler  evaporates  85  Ibs.  of  water  per  pound  of  coal.     Pressure  in  % 
boiler,  125  Ibs.;  feed  temperature,  150°.f    What  will  it  evaporate  from  and 
at  212°? 

4.  A  boiler  evaporates  8  Ibs.  of  water  per  pound  of  coal.     Pressure,  100 
Ibs.;  feed  temperature,  100°.     What  will  it  evaporate  if  the  pressure  is  80 
Ibs.  and  feed  200°;  and  what  will  it  evaporate  from  and  at  212°? 

5.  A  boiler  evaporates  8  Ibs.  of  water  per  pound  of  coal.     Steam  pressure, 
120  Ibs. ;  feed  temperature,  150°.     What  will  it  evaporate  with  a  steam  pres- 
sure of  5  Ibs.  and  a  feed  temperature  of  200°? 

/  6.  A  boiler  evaporates  9  Ibs.  of  water  per  pound  of  coal.  Steam  pres- 
r^/sure,  100 Ibs.;  feed  temperature,  50°.  What  will  it  evaporate  if  the  steam 
/  \pressure  is  200  Ibs.  and  the  feed  temperature  150°? 

7.  A  boiler  evaporates  8000  Ibs.  of  water  per  hour.      Steam  pressure, 
120  Ibs.;  feed  temperature,  180°.     What  would  it  evaporate  if  the  steam 
pressure  were  60  Ibs.  and  the  feed  temperature  60°? 

8.  A  boiler  plant  evaporates  6  Ibs.  of  water  per  pound  of  coal.     Steam  - 
pressure,  150  Ibs.;  feed  temperature,  120°.     What  will  it  evaporate  if  an 
economizer  is  added  increasing  the  feed  temperature  to  230°? 

9.  A  boiler  evaporates  5000  Ibs.  of  water  per  hour  from  a  feed-water- 
temperature  of  70°  into  steam  at  120  Ibs.  pressure.     What  is  the  evaporation 
from  and  at  212°?     If  the  efficiency  of  the  boiler,  furnace  and  grate  is  70  per 
cent,  and  coal  that  contains  13,500  B.T.U.  per  pound  is  used,  how  many 
pounds  of  water  will  be  evaporated  from  and  at  212°  per  pound  of  coal? 

10.  A  coal  contains  14,000  B.T.U.  per  pound  dry.     If  all  the  heat  in  this 


134  HEAT  ENGINES 

coal  should  be  utilized,  how  many  pounds  of  water  would  be  evaporated  per 
pound  of  dry  coal?  Steam  pressure,  200  Ibs.;  feed  temperature,  250°. 
*"•  11.  Efficiency  of  a  boiler,  furnace  and  grate  is  65  per  cent.  Coal  burned 
contains  12,000  B.T.U.  per  pound.  Steam  pressure,  200  Ibs.;  feed  tempera- 
ture, 180°.  How  many  pounds  of  water  will  be  evaporated  per  pound  of 
coal? 

"*  12.  A  boiler  burns  coal  containing  13,000  B.T.U.  per  pound.  Steam  pres- 
sure, 200  Ibs.;  feed  temperature,  200°;  efficiency  of  the  boiler,  furnace  and 
grate,  75  per  cent.  What  would  be  evaporated  from  and  at  212°? 

13.  One  hundred  pounds  of  coal  containing  13,000  B.T.U.  per  pound 
will  evaporate  how  many  pounds  of  water  at  200°  into  steam  at  100  Ibs. 
pressure?  What  will  it  evaporate  from  and  at  212°?  Efficiency  of  the 
boiler,  furnace  and  grate,  70  per  cent. 

-*  14.  How  many  pounds  of  water  can  be  evaporated  from  and  at  212° 
by  the  heat  evolved  by  the  complete  combustion  of  1  Ib.  of  dry  coal  contain- 
ing C,  65.2  per  cent.;  H,  4.92  per  cent.;  O,  8.64  per  cent.? 
—  16.  A  coal  contains  C,  75  per  cent.;  H,  5  per  cent.;  O,  4  per  cent.  Effi- 
ciency of  the  boiler,  furnace  and  grate,  70  per  cent.;  feed  temperature,  180°; 
steam  pressure,  150  Ibs  absolute.  Steam  contains  2  per  cent,  moisture,  (a) 
What  is  the  actual  evaporation  per  pound  of  coal?  (6)  What  is  the  equiva- 
lent evaporation  from  and  at  212°  per  pound  of  coal? 

*•  16.  A  coal  contains  C,  80  per  cent. ;  O,  7  per  cent. ;  H,  3  per  cent. ;  and  ash, 
10  per  cent.  It  is  used  in  a  boiler  carrying  100  Ibs.  pressure  with  a  feed  tem- 
perature of  180°.  The  efficiency  of  the  boiler,  furnace  and  grate  is  70  per 
cent.  What  will  be  the  evaporation  per  pound  of  coal? 
m  17.  If  40  per  cent,  of  the  heat  of  combustion  of  coal  containing  12,750 
B.T.U.  per  pound  is  lost,  how  many  pounds  of  coal  will  be  required  .to 
evaporate  5650  pounds  of  water  from  an  initial  temperature  of  130°  and 
under  a  pressure  of  80  Ibs.? 

18.  A  coal  contains  12,500  B.T.U.  and  requires  24  Ibs.  of  air  per  pound 
to  burn  it.  Temperature  of  boiler  room,  70°;  temperature  of  stack  gases, 
500°.  What  per  cent,  of  the  heat  of  the  coal  goes  up  the  stack? 
'**  19.  If  the  temperature  of  the  boiler  room  is  70°  and  the  temperature  of  the 
stack  gases  is  500°  and  30  Ibs.  of  air  are  used  per  pound  of  coal,  what  per  cent, 
of  heat  is  lost  up  the  stack,  if  the  coal  contains  14,500  B.T.U.  per  pound? 
-  20.  A  boiler  evaporates  3500  Ibs.  of  water  per  hour  from  an  initial  tem- 
perature of  120°  and  under  a  pressure  of  80  Ibs.  A  second  boiler  evaporates 
4000  Ibs.  of  water  from  an  initial  temperature  of  110°  and  under  a  pressure 
of  60  Ibs.  Which  of  the  two  boilers  utilizes  the  greater  amount  of  heat  per 
hour? 

21.  A  boiler  evaporates  6000  Ibs.  of  water  per  hour.  Coal  contains 
13,000  B.T.U.  Steam  pressure,  100  Ibs.;  feed  temperature,  180°;  efficiency 
of  boiler  and  grate,  70  per,  cent.  How  many  pounds  of  coal  will  the  boiler 
burn  per  hour?  /\k 

""      22.  An  engine  uses  30  Ibs.   of  steam  per  I.H.P.  per  hour.     Feed  tem- 
.    perature,  120°;  steam  pressure,  120  Ibs.     The  boiler  evaporates  9  Ibs.  of 
water  per  pound  of  coal.     How  many  pounds  of  coal  are  required  per  I.H.P. 
per  hour? 

23.  A  boiler  evaporates  7.5  Ibs.  of  water  per  pound  of  coal.     Steam 


BOILER  AUXILIARIES  135 

pressure,  150  Ibs. ;  feed  temperature,  200°.     Coal  costs  $2.50  per  ton.     What' 
is  the  cost  to  evaporate  1000  Ibs.  of  water  from  and  at  212°? 
••   24.  A    72-in.  return  tubular  boiler  18  ft.  long  has  seventy  4-in.  tubes 
Find  the  heating  surface  and  rated  B.H.P.  (Boiler  Horse-power). 

25.  A  66-in.  boiler  16  ft.  long  has  ninety-eight  3-in.  tubes.  Find  the 
heating  surface  and  rated  B.H.P. 

—  26.  A  60-in.  boiler  16  ft.  long  has  forty-four  4-in.  tubes.     Find  the  heating  - 
surface  and  rated  B.H.P. 

27.  A  60-in.  boiler  16  ft.  long  has  fifty-six  3^-in.  tubes.     Find  the  heating 
surface  and  rated  B.H.P. 

28.  A  48-in.  boiler  12  ft.  long  has  twenty-six  4-in.  tubes.     Find  the  heat- 
ing surface  and  rated  B.H.P. 

29.  A  36-in.  boiler  12  ft.  long  has  twenty-six  3-in.  tubes.     Find  the  heating 
surface  and  rated  B.H.P. 

••   30.  A  boiler  evaporates  4000  Ibs.  of  water  per  hour  from  a  feed  tempera- 
ture of  60°  into  steam  at  150  Ibs.  pressure  and  100°  of  superheat.     What  is 
the  factor  of  evaporationT^oiler  H.P.,  and  number  of  pounds  of  coal  used  per<« 
hour,  if  the  boiler,  furnace  and  grates  combined  have  an  efficiency  of  70  per 
cent,  and  the  coal  contains  14,000  B.T.U.  per  pound  dry. 

31.  A  boiler  evaporates  9000  Ibs.  of  water  per  hour.  Steam  pressure, 
150  Ibs.;  feed  temperature,  120°.  How  many  boiler  horse-power  is  it 
developing? 

—  32.  What  is  the  H.P.  of  a  boiler  which  evaporates  3080  Ibs.  of  water  per 
hour  from  an  initial  temperature  of  135°,  ari£  under  a  pressure  of  100  Ibs.? 

•  33.  A  1000-H.P.  engine  uses  15  Ibs.  ef  steam  per  H.P.  per  hour.     Steam 
pressure  at  boiler,  180  Ibs.;  feed  water  temperature,  120°.     What  boiler  H.P. 
should  we  have  to  supply  steam  for  the  engine? 

34.  A  boiler  evaporates  4000  Ibs.  of  water  per  hour  at  100  Ibs.  pressure 
from  a  feed  temperature  of  120°.     Quality  of  steam,  98  per  cent.     What  is 
the  boiler  H.P.? 

35.  A  fire-tube  boiler  is  60  in.  X  16  ft.  and  has  fifty-four  4-in.  tubes.     If 
it  evaporates  3000  Ibs.  of  water  per  hour,  is  it  working  over  or  under  its 
rated  H.P.  and  how  much?     Steam  pressure,  100  Ibs.;   feed  temperature 
200° 

—  36.  A  return  fire-tube  boiler  is  60  in.  in  diameter,  16  ft.  long,  and  has 
fifty-two  4-in.  tubes.     It  evaporates  4000  Ibs.  of  water  per  hour.     Steam 
pressure,  100  Ibs.;  feed  temperature,  150°.     Is  it  working  above  or  below  its 
rated  H.P.,  and  how  much? 

37.  A  boiler  is  reported  to  evaporate  12.5  Ibs.  of  water  per  pound  of  coal. 
Coal  contains  13,000  B.T.U.  and  uses  24. Ibs.  of  air  per  pound  to  burn  it. 
Temperature  of  the  boiler  room,  70°,  and  of  the  stack,  550°.  Steam  pressure, 
100  Ibs.;  feed  temperature,  70°.  Would  this  result  be  possible?  If  not, 
how  many  pounds  of  water  could  the  boiler  evaporate  per  pound  of  coal? 

-  38.  A  plant  burns  1500  Ib.  of  coal  per  hour.     The  height  of  the  stack 
is  130  ft.     Temperature-  of  boiler  room  is  70°  and  of  the  stack  gases,  500°, 
and  24  Ibs.  of  air  are  used  to  burn  1  Ib.  of  coal.     Coal  contains  12,000  B.T.U. 
per  pound.     What  should  be  the  area  of  the  stack?     What  per  cent,  of  heat 
is  lost  up  the  stack?     What  is  the  pressure  of  the  draft  in  tenths  of  inches  of 
water? 


*   "* 


,  r. 


136  HEAT  ENGINES  /5.  H*» 

K/*rt     ~ 
39.  A  boiler  is  to  evaporate  12,000  Ibs.  of  water  per  hour.     Steam  pressure, 

100  Ibs.  ;  feed  temperature,  200°.  (a)  What  should  be  the  horse-power  of  the 
boiler?  (6)  How  many  square  feet  of  heating  surface  should  the  boiler  con- 
tain?  (c)  How  many  square  feet  of  grate  surface  should  it  have?  (d) 
What  should  be  the  area  of  the  breeching? 

—  40.  In  a  100  H.P.  boiler  plant  what  should  be  the  area  of  the  grates,  and 
the  diameter  of  the  stack,  if  the  stack  is  125  ft.  high?  If  the  plant  carries 
130  Ibs.  gage  pressure,  would  a  water  or  a  fire-tube  boiler  be  used,  and  why? 

—  41.  A  400  H.P.  Corliss  engine  uses  26  Ibs.  of  steam  per  H.P.  per  hour. 
The  auxiliaries  use  25  per  cent,  as  much  as  the  engine.     The  boilers  to  supply 
the  plant  should  contain  how  many  square  feet  of  heating  surface  and  grate 
surface,  and  about  what  should  be  the  area  of  the  fluet     Pressure,  150  Ibs.; 
feed  temperature,  200°.     How  many  pounds  of  coal  will  the  plant  burn  per 
hour  if  the  coal  contains  13,500  B.T.U.  per  pound  and  the  efficiency  of  the 
boiler,  furnace  and  grate  is  70  per  cent.  ? 

-  42.  A  boiler  evaporates  7  Ibs.  of  water  per  pound  of  coal.     Steam  pressure, 
100  Ibs.  ;  feed  temperature,  50°.     A  feed-water  heater  is  added  increasing  the 
feed-water  temperature  to  200°.     Heater  costs  $400.     Allowing  5  per  cenfo 
interest  and  5  per  cent,  for  depreciation  and  repairs,  would  it  pay  to  install 
the  heater  if  the  plant  burns  750  tons  of  coal  per  year,  coal  costing  $2.50  per 
ton? 

—  43.  A  boiler  evaporates  1Q  Ibs.  of  water  per  pound  of  dry  coal  from  and  at 
212°.     Dry  coal  contains  13,000  B.T.U.  pet  pound.     What  is  the  combined 
efficiency  of  the  boiler,  furnace  and  grate? 

-»  44.  A  boiler  evaporates  7.5  Ibs.  of  water  per  pound  of  coal.  Coal  con- 
tains: 13,000  B.T.U.  Steam  pressure,  100  Ibs.;  feed  temperature,  150°. 
What  is  the  combined  efficiency  of  the  boiler,  furnace  and  grate? 

—  45.  What  is  the  combined  efficiency  of  a  boiler,  furnace  and  grate  that 
evaporates  8  Ihs.  of  water  per  Ib.  of  coal  from  a  feed  temperature  of  150°  into 
steam  at  150  Ibs.  pressure?     Coal  contains  13,000  B.T.U.  per  pound. 

-  46.  A  boiler  evaporates  9  Ibs.  of  water  per  pound  of  dry  coal  containing 
13,500  B.T.U.  per  pound.     Steam  pressure,  100  Ibs.  ;  feed  temperature,  200°. 
What  is  the  combined  efficiency  of  the  boiler,  furnace  and  the  grate? 

•»  47.  A  coal  contains  C,  80  per  cent.;  H,  4  per  cent.;  O,  2  per  cent.  What 
is  the  heat  value  of  the  coal?  If  this  coal  is  used  in  a  boiler  carrying  100  Ibs. 
pressure  with  a  feed  temperature  of  190°  and  evaporates  8  Ibs.  of  water  per 
pound  of  coal,  what  is  the  combined  efficiency  of  the  boiler,  furnace  and 
grate? 


48.  A  boiler  evaporates  15,000  Ibs.  of  water  per  hour  into  steam  at  100  Ibs. 


pressure;  temperature  of  feed,  200°.  Nine  pounds  of  water  are  evaporated 
per  pound  of  dry  coal  containing  13,000  B.T.U.  per  pound,  (a)  What  is  the 
H.P.  developed  by  the  boiler?  (6)  What  is  the  combined  efficiency  of  the 
boiler,  furnace  and  grate? 

^  «  49.  A  boiler  evaporates  11  Ibs.  of  water  from  and  at  212°  F.  per  pound  of 
dry  coal  containing  14,000  B.T.U.  per  pound.  What  is  the  combined 
efficiency  of  the  boiler,  furnace  and  grate?  At  the  same  efficiency,  what  will 
it  evaporate  with  a  steam  pressure  of  200  Ibs.  and  feed  temperature  at  200°? 

>-J  "50.  A  boiler  uses  1  Ib.  of  dry  coal  containing  13,000  B.T.U.  to  evaporate 
6  Ibs.  of  water.  Steam  pressure,  100  Ibs.;  feed  temperature,  100°.  (a) 


BOILER  AUXILIARIES  137 

What  is  the  efficiency  of  the  boiler  plant?  (&)  What  will  be  the  efficiency 
of  the  plant  if  a  heater  is  added  which  heats  the  feed  to  200°  F.?  (c)  What 
will  be  the  evaporation  per  pound  of  coal  after  the  feed-water  heater  is 
installed? 

51.  A  boiler  evaporates  9  Ibs.  of  water  per  pound  of  coal  fired.     Feed 
temperature,  70°;  steam  pressure,  150  Ibs.     Coal  as  fired  contains  3  per  cent, 
moisture.     Dry  coal  contains  14,000  B.T.U.  per  pound  and  has  6  per  cent, 
ash  by  analysis.     Twelve  per  cent,  of  coal  fired  is  taken  from  the  ash  pit 
in  form  of  ash  and  refuse,     (a)  What  is  the  efficiency  of  the  boiler  and 
furnace?     (6)  What  is  the  efficiency  of  the  boiler,  furnace  and  grates 
combined? 

52.  A  boiler  plant  burns  coal  which  contains  C,  75  per  cent.;  H,  6  per 
cent.;  and  O,  8  per  cent.     Two-thirds  of  the  carbon  is  burned  to  CO2  and  the 
balance  to  CO.     The  evaporation  is  8  Ibs.  of  water  per  pound  of  coal.     Steam 
pressure,  100  Ibs.;  feed  temperature,  170°.     (a)  What  is  the  efficiency  of  the 
boiler  and  furnace?     (6)  What  is  the  efficiency  of  the  boiler,  furnace  and 
grates  combined? 

53.  A  boiler  evaporates  20,000  Ibs.  of  water  per  hour  from  a  feed  tem- 
perature of  180°  into  dry  saturated  steam  at  115  Ibs.  pressure  absolute. 
Dry  coal  contains  4  per  cent,  ash  by  analysis  and  13,000  B.T.U.  per  pound. 
Ten  per  cent,  ash  and  refuse  are  taken  from  the  ash  pit.     The  actual  evapora- 
tion per  pound  of  dry  coal  is  9  Ibs.     (a)  What  H.P.   is  being  developed 
by  the  boiler?     (6)  What  is  the  efficiency  of  the  boiler,  furnace  and  grates 
Combined?     (c)  What  is  the  efficiency  of  the  boiler  and  furnace  alone? 

if  54.  Given  the  following  data  from  a  boiler  test :  Duration  of  test,  24 
/hours;  total  amount  of  water  fed  to  boilers,  240,000  Ibs.;  total  weight 
of  dry  coal  fired,  -30,000  Ibs.;  total  weight  of  ash  and  refuse,  3000  Ibs.;  tem- 
perature of  feed  water,  180°  F.;  steam  pressure,  150  Ibs.  absolute;  quality 
of  steam,  98.5  per  cent.;  dry  coal  contains  13,000  B.T.U.  per  pound  and 
3  per  cent,  ash  by  analysis,  (a)  What  H.P.  is  the  boiler  developing?  (6) 
What  is  the  evaporation  from  and  at  212°  per  pound  of  dry  coal?  (c)  What 
is  the  combined  efficiency  of  the  boiler,  furnace  and  grates?  (d)  What  is 
the  efficiency  of  the  boiler  and  furnace  alone?  (e)  What  should  be  the 
heating  and  grate  surfaces  in  this  boiler? 

55.  A  boiler  received  10,000  Ibs.  of  water  per  hour  at  100°  F.     Steam 
pressure,  150  Ibs.  absolute;  quality  of  steam,  98|  per  cent.     Dry  coal  burned 
per  hour,  1250  Ibs.,  each  pound  containing  13,000  B.T.U.     Per  cent,  of  ash 
by  analysis,  3  per  cent.;  ash  and  refuse  taken  from  ash  pit  per  hour,  125  Ibs. 
Coal  costs    $3   per  ton.     Plant  runs    10   hours    a   day,  300   days   in   the 
year,     (a)  What  H.P.  is  the  boiler  developing?     (6)  What  is  the  efficiency  of 
the  boiler,  furnace  and  grates  combined?     (c)  What  is  the  efficiency  of  the 
boiler  and  furnace  alone?     (d)  If  the  interest  and  depreciation  is  10  per 
cent.,  how  much  could  you  pay  for  a  heater  that  would  increase  the  tem- 
perature of  the  feed  water  to  212°? 

56.  A  boiler  evaporated  9000  Ibs.  of  water  per  hour  from  a  feed  tempera- 
ture of  80°  into  steam  at  145.8  Ibs.  absolute.     Coal  contains  13,500  B.T.U. 
and  costs  $2.50  per  ton.     Efficiency  of  the  boiler,  furnace  and  grate,  70  per 
cent.     If  we  add  a  feed-water  heater  that  will  increase  the  temperature  to 


138  HEAT  ENGINES 

212°,  what  will  be  the  saving  in  coal  cost  per  year,  if  the  plant  operates  10 
hours  a  day,  300  days  in  the  year? 

67.  A  boiler  plant  evaporates  30,000  Ibs.  of  water  per  hour.     Feed  tem- 
perature, 70°;  steam  pressure,  150  Ibs.     The  evaporation  is  8  Ibs.  of  water 
per  pound  of  coal,  and  coal  costs  $2.50  a  ton.     If  a  feed- water  heater  is  in- 
stalled that  will  increase  the  temperture  of  the  feed  water  to  180°,  how  much 
money  will  be  saved  per  year  and  how  much  can  be  paid  for  the  heater  if  the 
interest  and  depreciation  are  10  percent.?     Plant  runs  10  hours  per  day, 
300  days  in  the  year. 

68.  A  feed-water  heater  increases  the  temperature  of  the  feed  from  100° 
to  200°.     Steam  pressure,  100  Ibs.     The  plant  evaporates  10,000,000  Ibs.  of 
steam  per  year.     The  cost  to  evaporate  1000  pounds  of  steam  without  the 
heater  is  15  cents.     What  can  we  afford  to  pay  for  a  heater  allowing  5  per 
cent,  interest  and  8  per  cent,  depreciation  and  repairs? 

59.  A  boiler  plant  develops  500  B.H.P.  and  uses  4  Ibs.  coal  per  H.P.  per 
hour.     Coal  contains  13,000  B.T.U.  per  Ib.     Steam  pressure,  150  Ibs.     Feed 
temperature,  120°.     A  feed-water  heater  is  added  raising  the  temperature 
of  water  from  120°  to  200°.     Heater  costs  $500.     The  plant  operates   10 
hours  a  day  for  300  days  a  year.     The  cost  of  coal   is  $3  per  tori,    (a) 
Allowing  7  per  cent,  depreciation,  what  interest  will  the  owner  make  on 
the  investment?     (6)  If  later  an  economizer  is  added  which  raises  the  feed 
water  from  200°  to  300°,  allowing  5  per  cent,  interest  and  7  per  cent,  deprecia- 
tion, how  much  can  the  owner  pay  for  the  economizer?     (c)  What  would  be 
the  efficiency  of  the  plant  under  this  last  condition? 

60.  A  boiler  plant    runs  24  hours  per  day  for  300  days  in  the  year. 
It  burns  30  tons    of  coal   per    day    costing    $3    per    ton.     The    analysis 
of  the  stack  gases  is  CC>2,  5  per  cent.;  O,  15  per  cent.;  N,  75  per  cent.     The 
coal  contains  C,  80  per  cent. ;  H,  6  per  cent. ;  and  O,  4  per  cent.     The  plant  is 
changed  so  that  the  stack  gas  analysis  is  CO2,  14  per  cent.;  O,  6  per  cent.; 
N,  75  per  cent.     What  will  be  the  saving  in  dollars  per  year?     Stack  gas 
temperature,  600  F.     Boiler  room  temperature,  70°.     Boiler  radiation  loss, 
4  per  cent. 

After  this  change  is  made,  an  economizer  is  installed  which  reduces  the 
temperature  of  the  stack  gases  from  600°  to  400°.  The  evaporation  is  9  Ibs. 
of  water  per  pound  of  coal.  Feed-water  temperature  is  120°  P .  What  will 
be  the  final  temperature  of  the  feed  water?  What  will  be  the  saving  in 
dollars  per  year  after  this  second  change  is  made? 


CHAPTER  VIII 
STEAM  ENGINES 

91.  The  Simple  Steam  Engine. — A  simple  form  of  stationary 
steam  engine  and  one  in  general  use  is  shown  in  Fig.  54.  It  is  a 
small  direct  double-acting  engine  with  a  balanced  slide  valve 
and  a  cast-iron  cylinder  closed  at  its  ends  by  cylinder  heads 
bolted  on.  The  engine  has  no  steam  jacket  and  is  surrounded 
on  the  outside  by  non-conducting  material  and  cast-iron  lag- 
ging. Fig.  55  shows  the  steam  chest  containing  the  valves  and 
the  ports  leading  from  the  steam  chest  to  the  cylinder.  The 


FIG.  54. — Vertical  section  of  Skinner  engine. 

steam  is  admitted  and  exhausted  through  these  ports.  The 
piston  is  made  a  loose  fit  in  the  cylinder.  The  spring  rings 
shown  in  the  piston  serve  to  prevent  leakage  from  one  side  of 
the  piston  to  the  other.  The  piston  rod  is  usually  fastened  into 
the  piston  head  by  means  of  a  taper-ended  rod  and  nut,  and  is 
then  carried  through  the  cylinder  head,  the  gland  and  packing 
serving  to  make  a  steam-tight  joint.  The  other  end  of  the 
piston  rod  is  connected  with  the  cross-head.  The  power  is 

139 


140 


HEAT  ENGINES 


communicated  from  the  connecting  rod  to  the  crank,  which  is 
attached  to  the  main  shaft.  To  this  main  shaft  the  eccentric 
is  fastened  by  means  of  set-nuts.  The  valve  of  the  engine  is 
driven  by  the  eccentric  through  the  eccentric  rod  and  the  valve 
stem,  The  valve  stem  passes  through  the  steam  chest,  being 
made  tight  by  the  glands  and  packing,  as  in  the  case  of  the  piston 
rod,  and  is  fastened  by  lock  nuts  to  the  valve.  The  function 
of  this  valve  is  to  admit  the  steam  surrounding  the  valve  to 
each  end  of  the  cylinder  alternately.  On  the  opposite  stroke, 


Valve 


.Balance  Plate 
Steam  Port         /  Steam  Chf-st  Cover 


Packing 
Packing  Gland 


/      Packing 
Picking  Gland 


Lagging 

FIG.  55. — Section  through  steam  engine  cylinder  and  valve. 


Counter  Bore 
Piston 'Ring 


the  valve  opens  up  the  ends  of  the  cylinder  to  the  exhaust  space 
in  the  center  of  the  valve,  this  space  being  connected  to  the  ex- 
haust pipe  of  the  engine,  and  the  space  outside  of  the  valve  being 
connected  to  the  steam  pipe  admitting  the  steam  to  the  engine. 
Fig.  55  shows  the  slide  valve  in  a  position  admitting  steam  to 
the  head  end  of  the  cylinder.  On  the  crank  end,  the  cylinder  is 
open  to  exhaust.  As  the  steam  enters  behind  the  piston,  the 
steam  in  the  space  on  the  opposite  side  of  the  piston  is  forced  out 
through  the  space  under  the  valve  and  out  of  the  exhaust  port. 
When  the  piston  reaches  the  opposite  end  of  the  stroke,  the  valve 


STEAM  ENGINES 


141 


will  have  been  moved  to  a  similar  position  at  the  opposite  end. 
Steam  will  then  be  admitted  at  that  end,  and  the  end  previously 
receiving  steam  will  be  open  to  exhaust. 

92.  Action  of  the  Steam  in  the  Steam  Engine. — In  the  simplest 
form  of  steam  engine,  the  steam  is  admitted  for  the  full  stroke 
of  the  piston  and,  when  the  valve  opens  the  cylinder  to  exhaust, 
the  steam  is  exhausted  at  nearly  full  boiler  pressure.  This  action 
of  the  engine  is,  of  course,  very  uneconomical,  and  early  in  the 
development  of  the  engine  it  was  found  desirable  to  allow  the 
steam  to  expand  in  the  cylinder.  This  is  accomplished  by  having 
the  valve  close  the  entrance  port  before  the  piston  has  reached 
the  end  of  its  stroke,  then,  for  the  balance  of  the  stroke,  as  the 
piston  is  forced  out,  the  steam  pressure  in  the  cylinder  is  greatly 
reduced,  due  to  the  increased  volume  of  the  cylinder. 


FIG.  56. — Indicator  diagram. 

Fig.  56  shows  graphically  the  action  which  goes  on  in  the 
cylinder.  The  ordinates  of  the  diagram  represent  the  steam 
pressure,  and  the  abscissas  represent  cylinder  volumes  as  the 
piston  moves  out.  The  steam  enters  along  the  line  CDE,  the 
pressure  at  D  being  a  little  below  boiler  pressure.  At  the  point 
E,  known  as  the  point  of  cut-off,  the  valve  closes.  The  steam 
expands  from  the  point  E  to  F,  along  the  expansion  line  EF. 
At  the  point  F,  called  the  point  of  release,  the  valve  opens,  and 
from  the  point  F  to  the  point  H  the  exhaust  occurs.  At  the  point 
H  the  valve  closes  the  exhaust  port  and  compression  of  the  steam 
left  in  the  cylinder  begins,  continuing  along  the  line  HC  to  the 


142 


HEAT  ENGINES 


point  C.  At  this  point  steam  is  again  admitted  to  the  cylinder. 
A  similar  action  occurs  on  the  opposite  end  of  the  cylinder,  so 
while  the  steam  is  being  admitted  at  one  side,  at  the  opposite 
side  of  the  piston  we  have  exhaust  pressure.  Such  a  diagram  is 
termed  an  indicator  diagram  and  may  be  graphically  produced 
by  an  instrument  known  as  the  indicator. 

93.  Theoretical  Horse-power  of  a  Steam  Engine. — In  deter- 
mining the  theoretical  horse-power  of  a  steam  engine  it  is  as- 
sumed that  there  is  no  clearance,  that  the  full  pressure  of  steam 
is  maintained  during  admission,  that  the  cut-off  and  release  occur 
instantly,  and  that  the  engine  acts  without  compression.  Then 
the  indicator  card  of  the  engine  would  be  as  shown  in  Fig.  57. 


o  e 

FIG.  57. — Theoretical  indicator  card. 

The  curve  of  expansion  is  assumed  to  be  a  rectangular  hyperbola, 
the  equation  of  which  is  pv  =  a  constant,  as  this  is  the  curve 
which  coincides  most  nearly  with  the  actual  expansion  curve  in 
a  simple  non-condensing  engine. 

Let  the  pressure  at  the  point  of  cut-off  b  be  pi,  and  the  volume 
Vi]  and  let  the  pressure  at  the  point  d  be  pz,  and  the  volume,  vz. 
The  area  of  work  is  represented  by  the  area 
abode  =  oabg  +  gbcf  —  oedf. 


Area  oabg  = 


Area,  gbcf 


/v2 
Vi 


pdv.     Area  oedf  = 


Substituting  these  values  in  the  previous  equation,  the  area 
of  work, 

/l>2 
pdv  —  p2v2.  (1) 

»i 


STEAM  ENGINES  143 

t 

As  vi  and  ^  are  the  volumes  before  and  after  expansion,  fhe  ratio 
of  expansion, 


Since  the  expansion  curve  be  is  a  rectangular  hyperbola, 

pv  =  piVi. 

Hence  substituting  for  p  its  value  in  terms  of  pi  and  vi}  the 
equation  for  work  becomes 

/v2  dv  /         ("v2  dv\ 

-  -P&z   =  piVi  I    1  +    I  )    -  P&2. 

Vi  V  J    Vi   V  / 

Integrating,  and  substituting  r  for  --,  we  have 

abode  =  piVi(l  +  loger)  —  p2v2.  (3) 


The  average  pressure  on  the  card,  which  is  termed  the  mean 
effective  pressure,  is  found  by  dividing  this  by  the  length  of  the 
card  v2,  or 

M.E.P.  =  Pl-:  ±^-  -p*  (4) 

In  actual  practice,  however,  the  assumptions  made  are  not 
fulfilled,  and  the  actual  mean  effective  pressure  is  less  than  the 
theoretical  mean  effective  pressure.  The  proportion  borne  by 
the  actual  M.E.P.  to  the  theoretical  M.E.P.  is  termed  the  dia- 
gram factor,  e.  (Trans.  A.S.M.E.,  Vol.  24,  p.  751.) 

The  actual  mean  effective  pressure  is 

(5) 

This  diagram  factor  is  found  by  experiment  and  varies  from 
70  to  90  per  cent. 

To  determine  the  indicated  horse-power  of  a  steam  engine, 
it  is  necessary  to  find  the  work  done  in  the  engine  cylinder;  / 
Assume  the  engine  to  have  a  cylinder  a  square  inches  in  cross- 
section  and  1  ft.  long,  that  it  is  double-acting  and  makes  n 
revolutions  per  minute  (r.p.m.),  and  that  the  mean  effective 
pressure  determined  from  equation  (5)  acting  on  the  piston  is 
p  pounds  per  square  inch.  Then  the  total  pressure  against  the 
piston  will  be  pa  pounds  and  the  space  traveled  per  minute  by 
the  piston  will  be  2 In',  hence  the  foot-pounds  of  work  done  per 


144  HEAT  ENGINES 

minute  is  2  plan.     Since  1  horse-power  equals  33,000  ft.-lbs. 
per  minute,  the  indicated  horse-power  of  the  engine  is 

2plan 
:   33,000' 

Example. — A  12"  X  15"  double-acting  engine  runs  200  r.p.m.  Cut' 
oft7,  |  stroke;  steam  pressure,  100  Jbs.;  back  pressure,  2  Ibs.  absolute- 
Card  factor,  80  per  cent.  Find  the  rated  horse-power  of  the  engine. 

Solution. — From  equation  (2),  the  ratio  of  expansion, 

v2       1 

^  =  l  =  4'  \ 

and  from  equation  (5)  the 


M.E.P.  =  e<  ^(1  +  %-r) 

=  .80  [  i^p(l  +  log  A)  ~  2  j>    =  .80(28.7(1  +  1.39)  -  2) 
=  .80J68.5  -  2}  =  .80  X  66.5 
=  53 . 2  Ibs. 

The  cross-sectional  area  of  the  cylinder, 
a  =  ^2  =  3.U16  x  62 
=  113.3  sq.  in. 
From  equation  (6),  the 

I.H.P.  =  ?" 


33,000 
2  X  53.2  X  1.25  X  113.3'  X  200 

33,000 
#**-  91.4. 

Ans.  91.4  rated  horse-power. 

94.  Losses  in  a  Steam  Engine. — The  action  of  the  steam  in 
the  steam  engine  is  different  from  that  which  has  been  assumed 
as  the  ideal  action.  The  action  of  the  ideal  engine  is  useful, 
however,  as  a  basis  of  comparison  for  the  action  of  the  steam  in 
actual  engines.  In  the  actual  engine  the  steam  is  never  expanded 
completely,  and  has  at  the  end  of  the  expansion  a  higher  pressure 
than  the  back  pressure  in  the  exhaust  pipe.  It  is  not  advisable 
to  give  the  steam  complete  expansion,  as  there  will  be  no  added 
work  due  to  the  complete  expansion  of  this  steam,  the  pressure 
being  insufficient  to  overcome  the  friction  of  the  engine.  Qwing 


STEAM  ENGINES  145 

to  the  fact  that  we  do  not  have  complete  expansion,  it  is  necessary 
to  open  the  exhaust  valve  before  the  end  of  the  stroke  so  as  to 
bring  the  pressure  at  the  end  of  the  stroke  down  to  the  back 
pressure.  Comparing  the  ideal  diagram,  Fig.  57,  with  the  actual 
diagram,  Fig.  56,  it  will  be  noticed  that  the  steam  during  admis- 
sion in  the  actual  diagram  does  not  remain  at  full  boiler  pressure, 
but  that  there  is  a  reduction  of  the  pressure  due  to  wire  drawing 
of  the  steam  through  the  ports  of  the  valve.  In  the  ideal  engine 
there  is  no  transmission  of  heat  to  the  steam  except  in  the  boiler, 
but  in  the  actual  engine  there  is  a  transfer  of  heat  from  the  steam 
to  the  cylinder  walls  during  a  portion  of  the  stroke,  and  during 
other  portions  of  the  stroke  from  the  cylinder  walls  to  the  steam. 

In  an  actual  engine  the  back  pressure  in  the  cylinder  is  always 
greater  than  the  vacuum  in  the  condenser  owing  to  the  resistance 
of  exhaust  valve  and  passage.  In  the  ideal  engine  the  whole 
volume  of  the  cylinder  is  swept  through  by  the  piston,  and  in 
the  actual  engine  there  must  be  a  space  at  the  end  of  the  cylinder 
to  prevent  the  piston  striking  the  head. 

The  principal  losses  of  heat  from  an  engine  are  given  as  follows, 
as  nearly  as  possible  in  the  order  of  their  importance. 

1.  Heat  lost  in  the  exhaust.     This  loss  is  usually  70  per  cent, 
or  more  of  the  entire  heat  admitted  to  the  engine. 

2.  Initial  condensation. 

3.  Wire  drawing  at  admission  and  in  exhaust  valve. 

4.  Condensation  in  the  clearance  space  during  compression. 

5.  Radiation  and  conduction  from  the  cylinder. 

6.  Leakage  past  the  piston  and  valves. 

95.  Heat  Lost  in  the  Exhaust. — Most  of  the  heat  brought  to 
the  engine  by  the  steam  is  rejected  by  the  engine  in  the  exhaust. 
This  loss  varies  from  70  per  cent,  of  the  heat  of  the  steam  in 
the  best  engines  to  over  90  per  cent,  in  the  poorer  types.     In 
many  steam  plants  this  heat  is  partly  recovered  by  using  the 
exhaust  for  heating  or  manufacturing  purposes.     The  steam 
leaving  the  exhaust  of  an  engine  usually  contains  from  10_to 
20  per  cent,  of  water. 

96.  Initial    Condensation    and    Re -evaporation. — Early    ex- 
perimenters in  steam-engine  economy  found  that  the  surfaces  of 
the  cylinder  wall  and  steam  ports  played  a  very  important  part 
in  the  economy  of  the  steam  engine.     The  inner  surfaces  exposed 
to  the  action  of  the  steam  in  the  engine  naturally  have  a  tempera- 
ture very  close  to  that  of  the  steam  itself.     When  the  steam 

10 


146  HEAT  ENGINES 

enters  the  cylinder,  it  comes  in  contact  with  the  walls  of  the 
cylinder  which  have  just  been  exposed  to  exhaust  steam  and  are 
necessarily  at  a  lower  temperature.  A  part  of  this  steam  will, 
therefore,  be  condensed  in  warming  the  walls,  and  as  the  piston 
moves  out  more,  more  of  the  walls  will  be  exposed,  so  that  con- 
densation increases  to  a  point  even  beyond  the  point  of  cut-off. 
After  the  point  of  cut-off  the  steam  expands,  the  pressure  falls, 
and  the  temperature  drops  until  a  point  is  reached  where  the 
temperature  of  the  cylinder  walls  is  about  the  same  as  the  tem- 
perature of  the  steam  in  the  cylinder.  Condensation  ceases 
at  this  point  and  the  cylinder  walls  are  by  this  time  covered  with 
a  film  of  moisture.  If  the  expansion  of  the  steam  is  still  further 
increased,  the  temperature  in  the  cylinder  corresponding  to  the 
steam  pressure  will  be  less  than  the  temperature  of  the  cylinder 
walls,  and  this  film  of  moisture  on  the  surface  will  begin  to  re- 
evaporate.  Usually  the  amount  of  re-evaporation  is  very  much 
smaller  than  the  initial  condensation  and  the  cylinder  walls  are 
still  wet  when  the  exhaust  valves  open.  This  re-evaporation 
also  continues  during  the  exhaust  stroke.  It  is  very  desirable 
that  all  the  moisture  on  the  surface  of  the  cylinder  be  evaporated 
before  the  end  of  the  exhaust.  If  it  is  not  evaporated,  the  cylin- 
der walls  will  be  wet  when  steam  is  again  admitted  to  the  cylinder 
and  the  initial  condensation  will  be  greatly  increased.  The  trans- 
fer of  heat  from  the  steam  to  the  walls  of  the  cylinder  is  always 
increased  by  the  presence  of  moisture. 

In  the  average  non-condensing  engine,  initial  condensation  is 
from  15  to  20  per  cent.,  in  small  reciprocating  steam  pumps 
an  initial  condensation  as  high  as  75  per  cent,  sometimes  occurs, 
and  in  the  most  perfect  engines  it  is  from  10  to  12  per  cent. 

97.  Factors  Affecting  Initial  Condensation. — Initial  conden- 
sation is  always  increased  by  increasing  the  range  of  temperature 
in  the  cylinders. 

It  also  increases  as  the  proportion  of  the  area  of  the  cylinder 
walls  to  the  volume  of  the  cylinder  increases.  The  greater  this  ratio, 
the  less  the  economy,  as  the  more  wall  that  is  exposed  the  more 
heat  the  wall  will  take  up.  This  accounts  for  the  large  consump- 
tion of  steam  shown  by  most  rotary  engines. 

Time  is  also  important,  and  other  conditions  being  the  same, 
the  slower  the  speed  of  the  engine,  the  greater  the  initial  con- 
densation, as  the  whole  action  depends  upon  the  time  during 
which  the  heat  has  an  opportunity  to  be  taken  up  or  given  off  by 


STEAM  ENGINES  147 

the  cylinder  walls.  As  the  element  of  time  during  which  the 
steam  is  in  contact  with  the  walls  of  the  cylinder  increases,  the 
initial  condensation  increases.  The  changes  of  temperature 
only  affect  the  inner  surfaces  of  the  cylinder,  and  the  greater 
the  time,  the  greater  the  depth  of  cylinder  walls  that  will  be 
affected. 

Initial  condensation  increases  as  the  ratio  of  expansion  is 
increased,  that  is,  as  the  cut-off  becomes  shorter.  This  is 
easily  explained;  as  the  cut-off  is  shortened,  the  weight 
of  steam  admitted  to  the  cylinder  becomes  less  and  the 
amount  of  heat  taken  up  by  the  cylinder  walls  remains  sub- 
stantially the  same,  so  that  the  proportion  of  steam  condensed 
increases.  With  very  short  cut-offs  this  initial  condensation 
becomes  very  large.  When  the  cut-off  is  reduced  below  a  certain 
point,  the  increased  initial  condensation  offsets  the  increase  in 
economy  due  to  longer  expansion.  If  the  cut-off  is  shortened  to 
less  than  this  point,  the  steam  consumption  of  the  engine  will  be 
increased.  The  point  of  greatest  economy  in  most  single-cylinder 
engines  is  from  one-quarter  to  one-fifth  stroke.  In  an  engine 
having  a  short  cut-off  and  using  a  high  steam  pressure,  the  econ- 
omy may  often  be  increased  by  reducing  the  steam  pressure, 
thereby  increasing  the  cut-off. 

98.  Steam   Jacket. — The   action   of   initial    condensation   is 
increased  by  the  loss  of  heat  through  the  cylinder  wall  by  con- 
duction.    This  may  be  reduced  by  surrounding  the  cylinder 
with  steam  at  boiler  pressure.     Such  an  arrangement  is  called  a 
steam  jacket.     The  effect  of  the  steam  jacket  is  to  reduce  initial 
condensation  and  to  increase  the  re-evaporation.     The  steam 
used  by  the  steam  jacket  is  always  charged  to  the  engine  as  though 
it  had  been  used  in  the  cylinder.     Engines  with  jackets  show 
increased  economy,  particularly  when  operated  at  slow  speed. 
The  higher  the  speed  of  the  engine,  the  less  is  the  element  of  time 
during  which  the  jacket  can  affect  the  steam  in  the  cylinder  and 
the  less  effective  the  jacket  becomes.     In  cases  of  slow-speed 
engines  with  large  ratios  of  expansion,  the  use  of  the  jacket  will 
show  a  saving  of  from  10  to  20  per  cent. 

99.  Superheating. — Superheating  the  steam  previous  to  its 
admission  to  the  engine  is  used  as  a  means  of  reducing  initial 
condensation.     A  sufficient  amount  of  superheat  should  be  given 
to  the  steam  so  that  on  admission  of  steam  to  the  cylinder,  the 
cylinder  walls  take  up  this  superheat  instead  of  condensing  the 


148  HEAT  ENGINES 

steam.  The  effect  of  this  is  to  leave  the  cylinder  walls  entirely 
dry,  reducing  the  amount  of  heat  which  would  be  conducted  to  the 
walls,  as  dry  gas  is  one  of  the  best  non-conductors  of  heat.  The 
experiments  of  Professor  Gutermuth  show  that  with  sufficient 
superheat  the  economy  of  a  simple  non-condensing  engine  may  be 
made  to  equal  that  of  a  compound  condensing  engine. 

100.  Compound  Expansion. — By  increasing  the  steam  pressure 
and  using  a  longer  range  of  expansion,  the  range  of  temperatures 
in  the  cylinder  of  a  steam  engine  is  much  increased,  thereby 
increasing  the  initial  condensation.     In  order  to  reduce  the  range 
of  temperatures  in  the  cylinder,  it  has  been  found  more  economical 
partially  to  expand  the  steam  in  one  cylinder  and  then  exhaust  the 
steam  into  a  second  cylinder  in  which  the  expansion  is  completed. 
By  this  means  the  range  of  temperature  in  each  cylinder   is 
reduced  and  initial  condensation  reduced.     Compound  cylinders 
are  only  used  when  the  steam  pressure  is  sufficiently  high  so 
that  the  initial  condensation  would  be  excessive  if  the  steam 
were  expanded  in  one  cylinder.     With  steam  pressures  less  than 
100  Ibs.,  compound  engines  are  seldom  used.     It  is  not  necessary 
to  use  compound  engines  for  less  than  125  Ibs.  pressure  unless 
the  ratio  of  expansion  is  very  large. 

101.  Wire  Drawing. — The  resistance  offered  by  the  valves, 
#orts,  and  passages  lowers  the  pressure  of  the  steam  in  the  cylinder 
during  admission  and  raises  the  pressure  during  exhaust.     As  the 
valves  do  not  close  instantly  when  the  valve  nears  the  point  of 
closing,  or  cut-off,  the  pressure  is  reduced  owing  to  the  small 
port  opening.     This  is  shown  by  the  rounded  corners  at  cut-off 
and  release.     This  resistance  is  often  called  "throttling"  or  "wire 
drawing."     The  effect  of  this  throttling  of  the  steam  is  to  slightly 
dry  the  steam  and,  if  it  were  absolutely  dry  to  start  with,  there 
would  be  a  small  amount  of  superheating.     It  will  be  noticed  in 
the  indicator  diagram,  Fig.  56,  that  the  initial  line  DE  is  not  abso- 
lutely horizontal,  but  that  there  is  a  gradual  reduction  of  pres- 
sure from  D  to  E.     The  initial  pressure  line  is  always  lower 
than  the  boiler  pressure,  owing  to  the  resistance  of  the  passages 
between  the  boiler  and  the  cylinder. 

The  steam  in  passing  through  the  piping  leading  to  the  engine 
loses  a  certain  quantity  of  heat,  with  a  corresponding  condensa- 
tion. It  is  customary  to  place  a  separator  in  the  main  just  before 
it  reaches  the  engine  so  that  this  water  of  condensation  can  be 
removed  from  the  steam.  « 


STEAM  ENGINES 


149 


102.  Clearance  and  Compression. — In  order  that  the  piston 
may  not  strike  the  end  of  the  cylinder,  it  is  necessary  to  leave  a 
small  space  between  the  piston  and  the  cylinder  head.  In 
addition  there  is  always  some  space  in  the  steam  ports  between 
the  valve  and  the  cylinder.  The  volume  of  the  ports  between  the 
valves  and  the  cylinder,  together  with  the  space  between  the  pis- 
ton at  the  end  of  its  stroke  and  the  cylinder  head,  is  called  the 
clearance.  It  is  usually  determined  by  placing  the  piston  at  the 
extreme  end  of  its  stroke  and  filling  the  clearance  space  with 
water.  Knowing  the  weight  and  temperature  of  the  water  put 
into  the  clearance  space,  the  volume  of  the  water  may  be  deter- 
mined. Dividing  the  volume  of  the  clearance  by  the  volume  of 


O  G 

FIG.  58. — Theoretical  indicator  card  showing  clearance. 

the  piston  displacement  gives  the  per  cent,  of  clearance.  The 
clearance  in  engines  varies  from  1  to  10  per  cent.  The  steam  in 
the  clearance  affects  the  expansion  curve  of  the  engine. 

In  Fig.  58,  ED  represents  the  piston  displacement,  and  AB 
represents  the  volume  of  the  steam  admitted  to  the  cylinder. 
The  apparent  ratio  of  expansion  is 

ED. 

AB 

Actually,  however,  the  steam  expanding  includes  not  only  the 
steam  admitted  from  the  .boiler,  but  also  the  steam  left  in  the 
clearance,  so  that  the  real  ratio  of  expansion  is 

ED+AF 


AB+AF 


(8) 


150  HEAT  ENGINES 

The  clearance  of  the  engine  alters  the  amount  of  steam  con- 
sumed per  stroke  of  the  engine.  In  order  to  reduce  the  amount  of 
live  steam  to  fill  the  clearance  at  each  stroke,  the  exhaust  valves 
of  the  engine  are  closed  before  the  end  of  the  stroke,  and  for  the 
balance  of  the  stroke  the  steam  is  compressed.  This  compression 
of  the  steam  serves  to  fill  the  clearance  space  with  steam  at  a 
higher  pressure  than  the  exhaust  pressure.  In  addition,  compres- 
sion of  steam  in  the  clearance  space  serves  to  retard  the  reciprocat- 
ing masses  in  the  engine  and  bring  them  to  rest  at  the  end  of  the 
stroke.  If  an  engine  is  operated  with^too  little  compression,  it 
will  be  found  to  pound  at  the  end  of  the  stroke.  The  effect  of 
compression,  or  the  cushioning  of  the  piston,  is  materially 
increased  by  the  lead  of  the  engine.  The  lead  is  the  amount 
the  valve  is  open  when  the  piston  reaches  the  end  of  its  stroke.  In 
order  to  have  lead  it  is  necessary  to  open  the  valves  before  the  end 
of  the  stroke,  and  this  steam  admitted  before  the  end  of  the  stroke 
serves  to  assist  in  cushioning  the  piston  and  reciprocating  parts. 

We  may  consider  the  steam  occupying  the  cylinder  as  composed 
of  two  parts:  the  part  that  has  been  left  in  the  clearance,  which 
is  called  cushion  steam;  and  the  part  that  has  been  admitted  from 
the  boiler,  which  is  called  cylinder  feed.  If  it  is  desired  to 
determine  the  amount  of  steam  that  is  expanding  in  an  engine, 
it  is  necessary  to  add  to  the  weight  of  the  steam  fed  from  the  boiler 
the  weight  of  the  steam  left  in  the  clearance  space.  The  sum  will 
be  the  total  steam  expanding  in  the  engine. 

The  compression  of  the  steam  in  the  clearance  space  always 
involves  a  loss.  Just  previous  to  compression,  the  cylinder  walls 
have  been  exposed  to  the  exhaust  steam.  During  compression 
the  steam  compressed  has  its  temperature  increased,  and  when 
tlie  temperature  of  the  compressed  steam  exceeds  the  temperature 
of  the  walls,  condensation  begins  to  occur.  The  action  is  similar 
to  initial  condensation. 

PROBLEMS 

1.  An  electrical  plant  runs  a  factory  having  five  10  H.P.  motors,  two 
20  H.P.  motors,  four  30  H.P.  motors.     Efficiency  of  the  motors,  80  per  cent.  ; 
transmission,  80  per  cent. ;  of  the  engine  and  dynamo  combined,  80  per  cent. 
What  should  be  the  H.P.  of  the  engine  plant  and  kilowatts  of  the  generator? 

2.  A  street  car  plant  uses  ten  cars  each  requiring  an  average  horse-power 
of  75  at  the  wheels.     Efficiency  of  car  is  60  per  cent.;  of  transmission,  75 
per  cent.;  of  sub-stations,  75  per  cent.;  and  of  main  engines  and  dynamo, 
75  per  cent.     M.E.P.  of  engine,  40  Ibs.;  r.p.m.,  150.     Plant  has  two  engines 


STEAM  ENGINES  151 

of  equal  size.     What  are  the  dimensions  of  their  cylinders?     Assume  600 
ft.  per  minute  piston  speed. 

3.  Assume  the  mean  effective  pressure  to  be  40  Ibs.,  the  number  of  revo- 
lutions to  be  75  per  minute,  and  the  length  of  the  stroke  to  be  42  in.,  and 
determine  the  diameter  of  the  cylinder  of  a  Double-acting  engine  which  will 
develop  200  H.P. 

4.  An  engine  is  18"  X  36"  and  runs  100  r.p.m.     Initial  pressure,  100 
Ibs.;  back  pressure,  atmospheric;    cut-off,  i  stroke.     What  H.P.  will  be 
developed?  Assume  card  factor  of  85  per  cent. 

6.  An  engine  is  8"  X  12";  initial  steam  pressure,  100  Ibs.  gage;  back 
pressure,  3  Ibs.  gage;  cut-off,  £;  the  expansion  curve  is  an  isothermal  of  a 
perfect  gas;  r.p.m.,  250.  What  is  the  horse-power  of  the  engine?  Card 
factor,  85  per  cent. 

6.  Determine  the  horse-power  of  a  13"  X  18"  double-acting  engine  when 
making  220  r.p.m.  while  taking  steam  at  80  Ibs.  gage  and  cutting  off  at  1 
stroke.     Neglect  the  clearance  and  assume  that  the  mean  back  pressure  is 
20.5  Ibs.  absolute,  and  that  the  card  factor  is  80  per  cent. 

7.  An  engine  is  18"  X  30";  cut-off,  |  stroke.     It  runs  100  r.p.m.     Initial 
steam  pressure,  80  Ibs.     Exhaust,  atmospheric.     What  would  be  the  increase 
of  horse-power  if  the  cut-off  was  increased  to  \  stroke  and  initial  pressure 
to  150  Ibs.?     Card  factor,  80  per  cent. 

^8.  An  engine  is  8"  X  !§."  and  makes  300  r.p.m.;  cut-off,  i  stroke;  ex- 
haust, atmosphere.  What  would  be  the  horse-power  of  the  engine  at  the 
following  gage  pressures:  60,  80,  100,  and  120  Ibs.?  Card  factor,  75  per 
cent. 

9.  What  would  be  the  horse-power  developed  under  the  different  condi- 
tions stated  in  Problem  8,  if  a  condenser  were  added  and  the  back  pressure 
reduced  to  2  Ibs.  absolute? 

10.  An  engine  is  18"  X  30";  runs  100  r.p.m.,  and. initial  pressure  is  100 
Ibs.     Atmospheric  exhaust.     A  condenser  is  added  bringing  the  exhaust 
down  to  2  Ibs.  absolute.     In  both  cases  cut-off  occurs  at  \  stroke.     Card 
factor,  80  per  cent,   (a)  How  much  is  the  horse-power  of  the  engine  in- 
creased? (6)     If  the  power  is  sold  for  $60  per  horse-power  per  year,  how 
mifch  could  be  paid  for  a  condenser,   allowing  5  per  cent,  for  interest  and 
6  per  cent  for  depreciation? 

11.  An  engine  is  to  develop  600  H.P.  at  a  piston  speed  of  600  ft.  per 
minute.     Initial  steam  pressure,   100  Ibs.     Exhaust  pressure,  1  Ib.  gage. 
Cut-off  at  \  stroke.     Speed,  100  r.p.m.     Card  factor,  85  per   cent,     (a) 
What  should  be  the  stroke  and  diameter  of  the  cylinder?     (6)  What  should 
be  the  diameter  if  the  back  pressure  is  2  Ibs.  absolute? 

J2.  An  engine  is  to  develop  1000  H.P.  at  \  cut-off  and  120  r.p.m.  Initial 
steam  pressure,  125  Ibs;  back  pressure,  atmospheric;  piston  speed  not  to 
exceed  720  ft.  per  minute.  What  should  be  the  dimensions  of  the  cylinder? 
Card  factor,  70  per  cent. 

13.  The  cylinders  of  a  locomotive  are  19  in.  in  diameter  and  have  a  24-in 
stroke.  The  driving  wheels  are  7  ft.  in  diameter,  and  the  mean  back  pres- 
sure against  which  the  piston  works  is  19  Ibs.  absolute.  Determine  the  horse- 
power developed  by  the  locomotive  when  taking  steam  at  150  Ibs.  gage 


152  HEAT  ENGINES 

and  cutting  off  at  f  stroke,  while  traveling  at  a  speed  of  40  miles  per  hour 
Card  factor,  75  per  cent. 

1,4.  An  engine  has  a  clearance  volume  which  is  0.08  of  the  volume  swept 
through  by  the  piston  per  stroke.  If  the  steam  be  cut  off  at  I  stroke,  what 
will  be  the  number  of  times  it  is  expanded? 

16.  A  12"  X  14"  double-acting  engine  develops  97  H.P.  when  running 
260  r.p.m.  and  at  f  cut-off.  Pressure,  70  Ibs.  What  is  the  weight  of  steam 
actually  used  per  I. H.P.  per  hour,  assuming  that  one-quarter  of  that  theo- 
retically required  is  lost  through  condensation,  radiation,  etc. 

16.  A  tank  contains  1000  cu.  ft.  of  air  at  a  pressure  of  1000  Ibs.  per  square 
inch  absolute  and  a  temperature  of  60°  F.     This  tank  is  used  to  run  an 
8"  X  12"  double  acting  air  engine;  \  cut-off;  200  r.p.m.     The  initial  pressure 
of  air  entering  the  engine  is  60  Ibs.  per  square  inch  absolute.     How  long 
will  the  tank  run  the  engine? 

17.  A  tank  contains  200  cu.  ft.  of  air  at  200  Ibs.  absolute  and  a  temperature 
of  60°  F.     How  long  will  it  operate  a  4"  X  6"  double-acting  air  engine  run- 
ning 100  r.p.m.?     Cut-off  |  stroke.     Engine  takes  air  at  60  Ibs.  absolute. 
Temperature  constant. 

18.  A  double-acting  compressed  air  locomotive  has  two  air  tanks  each 
3'  X  12'.     These  tanks  supply  two  8"  X  12"  cylinders.     The  cylinders  take 
their  air  through  a  pressure  reducing  valve  at  100  Ibs.  per  square  inch  abso- 
lute, the  original  pressure  in  the  tanks  being  1000  Ibs.  per  square  inch  abso- 
lute,    (a)  If  the  air  acts  at  a  constant  temperature  of  60°  F.  and  the  expan- 
sion in  the  engine  is  isothermal,  how  long  will  the  tanks  run  the  engine  at  | 
cut-off  in  the  cylinder?     (6)  How  many  horse-power  will  be  developed  when 
the  engine  runs  150  r.p.m.,  assuming  a  card  factor  of  90  per  .cent.? 


CHAPTER  IX 
TYPES  AND  DETAILS  OF  STEAM  ENGINES 

103.  Classification. — Engines  may  be  classified  according  to 
whether  they  exhaust  into  the  atmosphere  or  into  a  condenser, 
into: 

1.  Non-condensing  engines. 

2.  Condensing  engines. 

In  the  non-condensing  engine  the  exhaust  passes  directly 
to  the  atmosphere.  In  condensing  engines  the  exhaust  steam 
passes  into  a  cold  chamber  where,  by  means  of  a  cooling  medium, 
the  steam  is  changed  to  water.  This  produces  a  vacuum  so 
that  the  exhaust  occurs  at  a  pressure  lower  than  that  of  the 
atmosphere.  The  condensed  steam  is  removed  and  the  vacuum 
is  sustained  by  means  of  an  air  pump. 

Another  classification  may  be  made  according  to  the  way  in 
which  their  speed  is  governed,  as: 

1.  Throttling  engines. 

2.  Automatic  engines. 

In  the  throttling  engines  the  speed  of  the  engine  is  controlled 
by  means  of  a  valve  in  the  steam  pipe  which  regulates  the  pressure 
of  the  steam  entering  the  engine.  In  the  automatic  engine  the 
pressure  of  the  entering  steam  remains  constant  and  the  governor 
controls  the  amount  of  steam  admitted  to  the  cylinder. 

Engines  may  also  be  classified  according  to  the  number  of 
cylinders  in  which  the  steam  is  allowed  to  expand  successively  as: 

1.  Simple  engines. 

2.  Compound  engines. 

3.  Triple  expansion  engines. 

4.  Quadruple  expansion  engines. 

In  a  simple  engine  the  steam  expands  in  but  one  cylinder, 
and  is  then  allowed  to  exhaust.  In  a  compound  engine  a  portion 
of  the  expansion  occurs  in  the  high-pressure  cylinder,  and  from 
there  the  steam  passes  to  the  low-pressure  cylinder,  where  it  is 
further  expanded  to  a  pressure  approximating  the  exhaust 
pressure.  In  the  triple-expansion  engine  the  steam  expands 
successively  in  three  cylinders,  and  in  the  quadruple  in  four. 

153 


154 


HEAT  ENGINES 


TYPES  AND  DETAILS  OF  STEAM  ENGINES      155 

A  fourth  classification  depends  upon  the  position  of  the  cyl- 
inder, as: 

1.  Vertical  engines. 

2.  Horizontal  engines. 

104.  Plain  Slide  Valve  Engine. — The  simplest  form  of  engine 
is  the  plain  D-slide  valve  engine,  as  shown  in  Fig.  59. 

The  valve  is  shown  in  its  normal  position  in  the  steam  chest. 
A  cross-section  of  a  valve  of  this  type  showing  the  steam  ports 
is  shown  in  Fig.  90. 


FIG.  60. — Portable  engine  and  boiler. 


This  type  of  engine  is  used  where  high  economy  is  not  neces- 
sary. It  requires  little  attention,  and  is  easily  repaired  and 
adjusted.  Fig.  60  shows  a  boiler  and  engine  of  this  type  arranged 
so  as  to  be  portable.  These  engines  are  governed  by  a  throttling 
governor  of  the  fly-ball  type,  as  shown  in  Fig.  60,  which  controls 
the  speed  of  the  engine  by  changing  the  pressure  of  the  steam  in 
the  steam  chest. 


156  HEAT  ENGINES 

105.  Automatic  High-speed  Engine. — This  class  of  engines  has 
developed  rapidly  since  the  introduction  of  electrical  lighting 
machinery,  and  is  designed  primarily  for  the  direct  driving  of 


FIG.  61. — Governor,  eccentric  rod,  rocker  shaft,  valve  and  valve  stem. 

electric  generators.  These  engines  have  balanced  slide  valves 
such  as  are  shown  in  Fig.  55.  The  governors  in  this  class  of 
engines  control  the  valve  directly,  and  it  is  necessary  that  the 


FIG.  62. — Bed  of  high-speed,  center-crank  engine. 

valve  be  balanced  so  that  it  may  be  moved  easily  by  the  governor. 
Fig.  61  shows  the  governor,  eccentric  rod,  rocker  shaft,  valve 
stem,  and  valve. 


TYPES  AND  DETAILS  OF  STEAM  ENGINES     157 


158 


HEAT  ENGINES 


TYPES  AND  DETAILS  OF  STEAM  ENGINES     159 

Engines  of  this  class  are  well  adapted  to  a  high  rotative  speed. 
The  stroke  of  these  engines  is  usually  short,  so  that  the  average 
piston  speed  may  exceed  600  ft.  per  minute  when  the  engine 
runs  at  a  large  number  of  revolutions  per  minute. 

Most  engines  of  this  class  are  of  the  center  crank  type  so  that 
all  parts  of  the  engine  are  supported  on  one  casting. 

Fig.  62  shows  the  bed  of  a  center-crank  high-speed  engine. 
This  bed  is  so  designed  that  all  parts  are  accessible  and  may  be 
removed.  It  may  be  machined  at  one  setting.  This  insures 
perfect  alignment  of  the  various  parts  of  the  engine.  This  bed 
casting  is  bolted  to  a  suitable  brick  or  cement  foundation. 

106.  Corliss  Engine. — These  engines  are  described  and  the 
action  of  their  valves  explained  in  paragraph  136.     Figs.  63  and  64 
show  a  plan  and  side  elevation  of  a  Corliss  engine. 

107.  The  Stumpf  Uniflow  Engine.— In  1910  Professor  Stumpf 
of  Charlottenburg,  Germany,  brought  out  an  engine  which  he 
called  a  uniflow  engine  and  which  promises  to  give  materially 
better  economy  than  the  ordinary  reciprocating  engine.     This 
engine  was  not  new  in  principle  as  the  patent  had  been  taken 
out  in  1886  but  had  not  been  used.     It  obtains  its  economy 
largely  through  reducing  the  initial  condensation  losses.     It  is 
of  the  four-valve  type. 

Fig.  65*  shows  a  section  through  the  cylinder.  This  cylinder 
has  no  exhaust  valves  but  in  the  middle  of  the  cylinder  there  is  a 
ring  of  ports  which  are  uncovered  by  the  piston  at  the  end  of  each 
stroke  so  that  the  piston  is  the  exhaust  valve.  There  are  two  steam 
valves  A  in  the  cylinder  heads,  and  the  steam  spaces  over  the  valves 
have  the  clearance  pockets  B  which  completely  steam  jacket  the 
heads.  In  the  uniflow  engine,  the  piston  faces  and  cylinder  heads 
are  exposed  to  exhaust  temperature  only  during  the  short  time 
that  the  piston  uncovers  and  covers  the  exhaust  ports.  On  the 
return  stroke  the  steam  remaining  in  the  cylinder  is  compressed 
in  the  clearance  spaces  up  to  the  admission  pressure.  The 
temperature  also  increases  in  the  compression  space,  not  only  due 
to  compression  but  also  to  the  absorption  of  heat  from  the 
jacketed  head. 

The  cylinder  Fig.  65  is  a  simple  cylindrical  casting  with  a  belt 
cast  in  the  middle  for  the  exhaust  passage.  The  steam  chest  is 
integral  with  the  cylinder  and  provided  with  two  drums  C  to  take 
up  expansion  without  distorting  the  cylinder.  Each  cylinder_has 

*  Taken  from  Power,  June  11,  1912,  vol  35,  No.  24,  p.  830. 


160 


HEAT  ENGINES 


a  large  valve  D  open  into  a  pocket  D  in  the  cylinder  head.  This 
valve  opens  automatically  to  serve  as  a  relief  valve  to  let  out 
entrained  water.  It  also  serves  as  extra  clearance  to  prevent  exces- 
sive pressure  if  the  vacuum  should  be  lost  when  the  engine  is 
operating  non-condensing.  In  the  particular  form  of  inflow 
engine  described  the  valves  are  Corliss  valves  and  operated  by 
the  usual  Corliss  valve  mechanism.  These  engines  have  shown 
very  low  steam  consumption  particularly  with  superheated 
steam.  In  a  recent  test  a  simple  single-cylinder  engine,  non- 
condensing,  developed  a  horse-power  with  11J  Ibs.  of  steam. 
In  addition  they  have  a  flat  economy  curve  and  are  capable 
of  taking  very  heavy  overloads. 


FIG.  65-. — Section  of  Stumpf  engine. 

108.  Engine  Details. — Fig.  66  shows  the  piston  and  piston  rod. 
The  piston  is  turned  a  little  smaller  than  the  cylinder,  and  is 
made  tight  in  the  cylinder  by  means  of  spring  rings.  These 
rings  are  shown  in  the  figure  leaning  against  the  piston  rod. 
They  are  made  of  cast  iron  and  are  so  constructed  that  they 
have  to  be  compressed  in  order  to  get  them  into  the  cylinder, 
and  when  the  piston  is  in  place,  the  rings  bear  firmly  against 
the  cylinder  walls.  The  piston  with  rings  in  place  is  shown  in 
Fig.  67.  In  Fig.  68  is  shown  a  piston,  piston  rod,  and  cross-head. 
The  piston  is  attached  to  the  piston  rod  by  a  taper  pin  and  lock- 
nut,  and  the  other  end  of  the  piston  rod  is  screwed  into  the  cross- 
head  and  fastened  by  a  lock-nut.  The  cross-head  pin  is  also 
shown  in  the  cross-head. 


TYPES  AND  DETAILS  OF  STEAM  ENGINES     161 

Fig.  69  shows  a  solid-ended  connecting  rod.  These  rods  are 
usually  made  of  forged  steel.  The  bearings  that  enclose  the  pin 
are  made  of  brass  and  fitted  into  the  ends  of  the  rods.  These 


FIG.  66. — Piston  and  piston  rings. 


FIG.  67. — Piston  with  rings  in  place. 


FIG.  68. — Piston,  piston  rod  and  crosshead. 

bearings,  or  brasses,  are  taken  up  when  they  wear  by  means  of 
wedges  held  by  lock-nuts  as  shown  in  the  cut. 

Fig.  70  shows  a  strap-ended  connecting  rod.     In  this  form  of 
rod  the  brasses  are  held  in  place  by  steel  straps  that  encircle 
them.     These  straps  are  fastened  to  the  body  of  the  connecting 
11 


162 


HEAT  ENGINES 


rod  by  means  of  a  taper  key  and  a  cotter.  The  brasses  in  this 
rod  are  shown  lined  with  babbitt  metal  which  is  much  softer 
than  the  steel  pins  themselves. 


FIG.  69. — Solid  end  connecting  rod. 

Fig.  71  shows  the  crank-shaft  and  its  counterbalance  weights 
which  are  bolted  to  the  crank.  The  crank-shaft  is  a  solid  forging 
of  open-hearth  steel.  The  counterweights  are  made  of  cast  iron. 


'•*!•' 


FIG.  70. — Strap  end  connecting  rod. 

The  crank-shaft  shown  in  the  figure  is  designed  for  a  center-crank 
engine. 

Fig.  72  shows  one  of  the  main  bearings  for  the  crank-shaft. 
The   figure   shows    what   is   called   a   four-part   bearing.     The 


FIG.  71. — Counter-balanced  crank. 

bearing  proper  is  made  up  of  four  pieces.     The  two  side  pieces, 
or  brasses,  take  up  most  of  the  wear  in  the  bearing  and  are 


TYPES  AND  DETAILS  OF  STEAM  ENGINES     163 


adjusted  by  means  of  set  screws  fastened  with  lock-nuts.  The 
upper  part  of  the  brasses  is  adjusted  by  a  screw  in  the  top  of  the 
bearing.  The  brasses  are  supported  by  the  main  frame  of 


FIG.  72. — Main  bearing,  four  part. 

the  engine  and  held  down  by  a  main  bearing  cap  bolted  to  the 
main  frame  of  the  engine. 

Fig.  73  shows  the  eccentric  strap  and  eccentric  rod.     The 
eccentric  strap  is  driven  by  an  eccentric  sheave  the  position  of 


FIG.  73. — Eccentric  strap  and  eccentric  rod. 

which  is  determined  by  the  governor.  Fig.  74  shows  the  eccen- 
tric sheave. 

In  Fig.  75  is  shown  the  eccentric  strap  more  in  detail.  The 
strap  is  split  in  two  parts  and  bolted  together  so  that  it  can  be 
placed  over  the  sheave. 

In  Fig.  76  is  shown  the  main  frame  for  a  side-crank  engine. 


164 


HEAT  ENGINES 


This  cut  shows  a  main  bearing  with  a  three-part  box.  The  side 
brasses  in  this  box  are  adjusted  by  wedges  moved  by  set-nuts 
on  the  top  of  the  bearing. 


FIG.  74. — Eccentric  sheave  for  shaft  governor. 

Figs.  77  and  78  show  two  views  of  a  main  engine-bearing 
having  an  oil  cellar.  The  lower  part  of  the  cellar  is  filled  with 
oil  which  is  carried  up  onto  the  bearing  by  means  of  a  chain 


FIG.  75.' — Eccentric  strap. 

which  hangs  over  the  shaft  and  dips  into  the  cellar.  The  chain 
is  moved  by  the  rotation  of  the  shaft,  bringing  the  oil  up  on  to 
the  shaft, 


TYPES  AND  DETAILS  OF  $TEAM  ENGINES     165 

109.  Lubricators. — Although  not  a  part  of  the  engine  itself, 
the  lubricator  is  so  intimately  associated  with  it  that  it  seems 
desirable  to  describe  its  action  at  this  point. 


FIG.  76. — Frame  for  side  crank  engine. 


FIG.  77. — Main  engine  bearing  with  oil  cellar — cross-section. 


A  cross-section  of  a  sight  feed  lubricator  is  shown  in  Fig.  79. 

The  lubricator  is  connected  to  the  steam  main  just  before  the 

main  enters  the  steam  chest,  and  its  purpose  is  to  supply  oil  for 


166 


HEAT  ENGINES 


lubricating  the  engine  cylinder.     This  oil  is  carried  into  the  cylin- 
der by  the  entering  steam. 


FIG.  78. — Main  bearing  with  oil  cellar — transverse-section. 


FIG.  79. — Sight-feed  lubricator. 


" Steam  being  admitted  into  pipe  'B'  and  condenser  'F'  con- 
denses, thus  forming  a  column  of  water  which  exerts  a  pressure 


TYPES  AND  DETAILS  OF  STEAM  ENGINES     167 

equal  to  its  head  plus  the  difference  in  specific  gravity  between 
oil  and  water,  through  the  tube  'P'  on  the  oil  in  reservoir  'A.' 
By  this  excess  pressure  the  oil  is  forced  from  reservoir  '  A'  through 
the  tube  'S'  and  sight  feed  nozzle  'N'  into  the  sight  feed  chamber 
1 H.'  The  sight  feed  chamber  being  filled  with  water,  the  drop 
of  oil  floats  to  the  top  and  passes  to  the  point  to  be  lubricated 
through  the  passage  "I"  and  support  arm  'K.'" 


CHAPTER  X 
TESTING  OF  STEAM  ENGINES 

110.  The  Indicator. — The  indicator  is  a  device  by  which  the 
pressure  of  the  steam  for  each  point  in  the  stroke  of  the  engine  is 
graphically  recorded.  It  was  first  invented  by  James  Watt 
and  has  since  reached  a  high  state  of  perfection. 

There  are  three  principal  things  determined  by  an  indicator: 


FIG.  80. — Crosby  indicator. 

First,  the  average  pressure  of  the  steam  acting  against  the 
piston,  which  is  called  the  mean  effective  pressure  (M.E.P.). 

Second,  the  distribution  of  the  steam  in  the  engine;  that  is, 
the  point  at  which  the  valves  of  the  engine  are  opened  and  closed. 
By  the  use  of  the  indicator  we  are  able  to  determine  whether  or 
not  the  engine  has  a  proper  distribution  of  steam. 

168 


TESTING  OF  STEAM  ENGINES 


169 


Third,  from  the  indicator  we  may  determine  the  actual  weight 
of  steam  which  is  being  worked  in  the  engine  cylinder.  The 
indicator  makes  possible  a  complete  analysis  of  the  action  of  the 
steam  engine. 

Fig.  80  shows  a  cross-section  of  a  Crosby  steam-engine  indi- 
cator. This  instrument  is  attached  to  the  engine  cylinder,  and 
the  space  under  the  piston  8  is  in  direct  communication  with  the 


FIG.  81. — Crosby  indicator  with  outside  spring. 

engine  cylinder.  The  pressure  of  the  steam  acts  agianst  the 
piston  8,  compressing  a  spring  above  it.  The  pressure  of  the 
steam  raises  an  arm  16,  and  the  attached  pencil  at  23.  The  drum 
24  is  covered  with  a  sheet  of  paper;  a  cord  passing  over  a  pulley 
34  is  attached  to  the  engine  cross-head  through  a  reducing  motion, 
so  that  with  each  stroke  of  the  engine  the  drum  makes  almost  a 
complete  revolution.  The  movement  of  the  drum  corresponds 


170  HEAT  ENGINES 

to  the  movement  of  the  piston,  and  the  upward  movement  of  the 
pencil  corresponds  to  the  pressure  in  the  cylinder.  We  have  a 
diagram,  therefore,  of  the  pressure  in  the  cylinder  for  each  point 
in  the  stroke  of  the  engine.  The  springs  used  above  the  piston 
are  of  various  strengths.  What  is  termed  a  40-lb.  spring  would 
be  one  of  such  strength  that  a  pressure  of  40  Ibs.  per  square  inch 
under  the  piston  would  move  the  pencil  one  inch.  These  springs 
are  carefully  calibrated  so  that  certain  movements  of  the  piston 
give  a  corresponding  movement  of  the  pencil  on  the  paper. 

A  brass  stylus  is  sometimes  used  in  place  of  a  pencil.  This  has 
the  advantage  of  always  keeping  a  sharp  point.  In  this  case  the 
indicator  cards  are  taken  on  a  specially  prepared  paper  with  a 
metallic  surface,  as  no  mark  would  be  made  on  ordinary  paper. 


Elevation.  Cross-section. 

FIG.  82. — Thompson  indicator. 

One  disadvantage  of  the  use  of  the  stylus  and  metallic  surfaced 
paper  is  that  the  outline  traced  by  the  brass  point  is  not 
permanent,  but  will  fade  out  in  a  comparatively  short  time. 

Fig.  81  shows  a  similar  indicator  with  the  spring  external  to 
the  indicator  piston.  The  temperature  of  the  spring  in  this 
indicator  is  independent  of  the  steam  pressure.  The  spring 
in  this  indicator  may  easily  be  changed  without  removing  the 
indicator  piston.  This  form  is  particularly  adapted  for  indicator 
work  where  great  accuracy  is  desired. 

Fig.  82  shows  the  elevation  and  cross-section  of  the  Thomp- 
son indicator.  This  form  of  indicator  is  particularly  well 
adapted  to  hard  service. 

111.  Use  of  Indicator. — The  accuracy  of  an  indicator  depends 
upon  the  accuracy  with  which  the  pressure  in  the  cylinder  is 


TESTING  OF  STEAM  ENGINES 


171 


recorded  on  the  indicator  drum,  and  also  upon  the  accuracy 
with  which  the  motion  of  the  piston  is  conveyed  to  the  indicator 
drum.  In  order  to  have  the  pressure  recorded  properly,  the 
following  conditions  should  be  observed:  the  piping  leading  to 
the  indicator  should  not  be  more  than  18  in.  long,  and  should 
be  \  in.  in  diameter;  the  indicator  should  never  be  connected 
to  a  pipe  through  which  a  current  of  steam  is  passing;  the  holes 
connecting  the  indicator  with  the  cylinder  should  be  drilled  into 
the  clearance  space  so  that  the  piston  will  not  cover  the  opening ; 
the  indicator  should,  if  possible,  be  placed  in  a  vertical  position. 
Where  great  accuracy  is  desired  the  indicator  spring_should 
be  calibrated  before  and  after  the  test. 


FIG.  83. — Reducing  motion,  showing  method  of  attachment. 


The  motion  of  the  drum  may  be  taken  from  any  part  of  the 
engine  which  has  the  same  relative  motion  as  the  engine  piston. 
The  movement  of  the  drum,  which  is  usually  taken  from  the  cross- 
head,  must  be  reduced  to  the  length  of  the  indicator  diagram  by 
some  form  of  mechanism  which  makes  the  reduced  motion  an 
exact  ratio  to  the  movement  of  the  engine  piston.  The  indicator 
drum  is  then  connected  with  this  reduced  motion  of  the  piston 
by  means  of  a  cord.  A  reducing  lever  and  segment  is  one  of  the 
commonest  means  used  to  accomplish  this  reduction.  There 
are  also  on  the  market  various  forms  of  reducing  wheels  which 
make  the  reduction  by  means  of  gearing  and  pulleys.  These 
reducing  motions  are  more  satisfactory  when  they  are  provided 
with  a  clutch  so  that  the  drum  may  be  disengaged  without 


172  HEAT  ENGINES- 

removing  the  cord  connection  from  the  reducing  motion  to  the 
engine. 

Fig.  83  shows  a  simple  form  of  reducing  motion  made  of 
hard  wood  splines  and  a  brass  segment.  It  is  better  to  use  a 
segment  of  a  circle  at  the  point  6,  so  that  db  is  the  same  distance 
for  every  point  of  the  stroke. 

Fig.  84  shows  a  reducing  wheel  having  a  clutch,  so  that  it  is 
not  necessary  to  disconnect  the  motion  from  the  cross-head  when 
the  paper  on  the  drum  is  replaced. 

Cord  that  has  been  stretched  should  be  used  on  the  indicator 

and  reducing  motion,  so  that  the 
give  of  the  cord  will  not  reduce  the 
length  of  the  card.  Wherever  very 
long  cords  are  found  necessary,  it 
is  better  to  replace  the  cord  with 
piano  wire. 

FIG.  84.-Reducing  wheel.  112-  Taking  an  Indicator  Card.- 

Before  attaching  the  indicator,  oil 
the  parts  of  the  mechanism  with  watch  oil  and  the  piston  with 
cylinder  oil.  Be  sure  the  piston  is  working  freely  in  the  cyl- 
inder. The  piston  should  drop  by  gravity  in  the  cylinder  when 
the  spring  is  removed.  The  pencil  should  have  a  smooth,  fine 
point.  Be  sure  there  is  no  lost  motion  in  the  instrument. 

The  reducing  motion  should  be  adjusted  so  that  the  length 
of  the  card  is  from  2^  to  3  in.  The  higher  the  speed,  the  shorter 
should  be  the  card.  The  tension  of  the  indicator  drum  spring 
should  be  just  sufficient  to  prevent  slackness  in  the  cord.  Before 
taking  a  card,  try  the  indicator  and  see  that  it  does  not  strike 
the  stops  at  either  end  of  the  stroke.  The  cord  should  run  to 
the  indicator  over  the  center  of  the  guide  pulleys.  Steam  should 
be  turned  on  the  indicator  a  few  moments  before  taking  the  card 
so  as  to  warm  up  the  instrument. 

113.  To  Find  the  Power  of  the  Engine. — The  piston  area  is  the 
cross-section  of  the  cylinder.  The  diameter  of  the  cylinder 
should  be  obtained  with  a  caliper  and  the  corresponding  area  is 
the  piston  area  a.  The  piston  area  is  not  the  same  at  bojh  ends 
of  the  stroke,  as  on  the  crank  end  the  area  of  the  piston  rod  must 
be  subtracted. 

The  travel  of  the  piston  in  feet  per  minute  for  each  end  of  the 
stroke  is  found  by  multiplying  the  length  of  the  stroke  by  the 
revolutions  of  the  crank-shaft  per  minute. 


TESTING  OF  STEAM  ENGINES 


173 


The  mean  effective  pressure  is  obtained  from  the  indicator 
card.  The  usual  method  is  to  measure  the  area  of  the  card  with 
an  instrument  called  a  plartimeter. 

In  Fig.  85  is  shown  a  standard  form  of  planimeter.  In  using 
a  planimeter  the  point  B  is  placed  on  a  point  on  the  indicator  card 
to  be  measured  and  the  vernier  E  set  at  zero.  The  point  B  is 
then  made  to  trace  the  card  in  a  clockwise  direction,  going  all 
around  the  card  and  returning  to  the  starting-point.  The  reading 
of  the  scale  on  the  rotating  wheel  C  will  then  show  the  number  of 
square  inches  enclosed  by  the  diagram.  Dividing  the  area  of  the 


FIG.  85. — Polar  planimeter. 

card  by  the  length  of  the  diagram  will  give  the  average  height  of 
the  card  in  inches,  and  this  multiplied  by  the  value  of  the  spring 
gives  the  mean  effective  pressure,  M.E.P.  The  M.E.P.  should  be 
determined  for  each  end  of  the  cylinder  separately. 

The  mean  ordinate  from  the  card  may  also  be  obtained  by 
dividing  the  card  into  ten  spaces  by  vertical  lines  drawn  equi- 
distant apart.  Then  measure  the  distance  from  the  back  pres- 
sure line  to  the  forward  pressure  line  at  the  center  of  each  space. 
The  average  of  these  lengths  will  be  approximately  the  mean 
ordinate. 


174 


HEAT  ENGINES 


Let  ph  be  the  mean  effective  pressure  for  the  head  end,  and  pc, 
for  the  crank  end;  ah,  the  cross-sectional  area  of  the  piston  in 
square  inches  for  the  head  end,  and  ac,  for  the  crank  end;  I,  the 
length  of  the  stroke  in  feet;  and  n,  the  number  of  revolutions  per 
minute.  Then  the  indicated  horse-power  will  be 


IHP 


(2) 


The  total  I.H.P.  of  the  engine  is  the  sum  of  the  I.H.P.  for  the 
head  end  and  the  crank  end'. 

114.  Indicator  Diagrams.  —  The  indicator  is  very  often  used 
to  determine  the  setting  of  the  valve  and  the  distribution  of  steam 


o  o 

FIG.  86. — Indicator  card  from  non-condensing  Corliss  engine. 

in  the  cylinder.  Fig.  86  shows  a  typical  indicator  card  from  a 
Corliss  engine  running  non-condensing.  AB  is  the  atmospheric 
line,  and  00' ,  the  line  of  absolute  vacuum,  or  zero  pressure  abso- 
lute. OY  is  the  line  of  no  volume  for  the  head  end,  and  O'Y'  for 
the  crank  end  of  the  cylinder.  The  horizontal  distance  between 
the  lines  OY  and  CD  represents  the  clearance  volume  for  the  head 
end  of  the  cylinder.  The  clearance  on  the  crank  end  is  similarly 
shown. 

115.  Graphical    Determination    of    Initial    Condensation.— 
Initial  condensation  may  be  determined  graphically  from  the 


TESTING  OF  STEAM  /ENGINES 


175 


indicator  card.  In  determining  the  amount  of  steam  working  in 
the  engine  cylinder,  the  amount  supplied  to  the  engine  per  stroke 
is  determined  by  either  weighing  the  water  entering  the  boiler, 
which  passes  over  as  steam  into  the  engine,  or  by  weighing  the 
steam  condensed  in  a  condenser  attached  to  the  exhaust  of  the 
engine. 

This  total  quantity  of  steam  used  by  the  engine  is  then  reduced 
to  the  amount  of  steam  used  per  stroke,  and  this  is  called  the 
cylinder  feed.  To  this  must  be  added  the  cushion  steam.  To 
determine  the  amount  of  cushion  steam,  an  average  indicator 
card  is  selected,  and  at  a  point  after  compression  has  begun  and 
it  is  certain  that  the  valve  is  closed,  the  pressure  is  measured  and 
the  volume  determined.  This  volume  must  include  the  volume 
of  the  clearance.  From  this  pressure  and  volume,  by  reference 


FIG.  87. — Indicator  card  and  saturation  curve,  showing  effect  of  initial 

condensation.  , 


to  the  steam  tables,  the  weight  of  the  cushion  steam  may  then 
be  calculated,  assuming  the  steam  to  be  saturated.  The  total 
steam  in  the  cylinder  during  expansion  is  then  found  by  adding 
this  cushion  steam  to  the  cylinder  feed.  A  curve  of  saturation 
for  this  total  quantity  of  steam  can  then  be  drawn  upon  the  indi- 
cator diagram,  and  this  curve  will  represent  at  each  point  of  the 
stroke  the  volume  of  steam  if  no  initial  condensation  has  occurred. 
Fig.  87  shows  a  saturation  curve  constructed  on  an  indicator 
card.  YR  represents  the  volume  of  the  steam  as  supplied  to  the 
engine  per  stroke,  or  in  other  words,  it  represents  the  volume  of 
the  total  steam  in  the  cylinder  at  boiler  pressure  if  all  the  steam 
entering  remained  steam.  The  curve  RS  represents  the  volume 


176  HEAT  ENGINES 

of  this  same  weight  of  steam  for  the  varying  pressures  of  expan- 
sion. The  difference  in  the  volume  between  this  theoretical 
expansion  line  and  the  actual  expansion  line  represents  the  loss 
in  volume  due  to  condensation.  The  percentage  of  initial  con- 

C1! 

densation  at  the  point  of  cut-off  would  be  77 J  and  at  any  other 

kl 
point,  such  as  k,  would  be    .,-• 

Example. — An  8"  X  12"  engine  runs  230  r.p.m.  and  uses  700  Ibs. 
steam  per  hour.  Steam  pressure,  100  Ibs.;  exhaust,  atmospheric;  clear- 
ance, 10  per  cent.;  scale  of  indicator  spring,  60  Ibs.  Find  the  total 
weight  of  steam  in  the  cylinder  durmg  expansion. 

Solution. — First  find  the  cylinder  feed,  or  amount  of  steam  supplied 
by  the  boiler  to  the  engine  per  stroke. 

Strokes  per  hour  =  230  X  2  X  60  =  27,600. 
Cylinder  feed  =  700  -f-  27,600  =  .02536  Ibs. 

To  find  the  amount  of  cushion  steam,  first  lay  off  from  u,  Fig.  87,  the 
distance  uO  equal  to  10  per  cent,  of  uv,  since  the  clearance  is  10  per 
cent,  and  -uv  represents  the  volume  of  the  cylinder.  If  the  length  uv 
of  the  card  is  2.9  in.,  the  total  length  Ov  is  3.2  in. 

The  volume  swept  through  by  the  piston  is3.1416X4X4X12  = 
602.4  cu.  in.  The  clearance  volume  is  then  60.2  cu.  in.,  and  the  total 
volume  662.6  cu.  in.  In  other  words,  each  inch  of  length  of  the  line 
Ov  represents  662.6  -f-  3.2  =  207  cu.  in. 

Now  take  a  point  on  the  compression  curve  after  the  exhaust  valve 
has  closed,  such  as  N.  The  ordinates  of  this  point  measured  from  the 
axes  OY  and  OX  are,  p  =  34.8  Ibs.  absolute,  and  v  =  124.2  cu.  in. 
=  .07187  cu.  ft.  From  the  steam  tables  we  find  that  1  cu.  ft.  of  dry 
saturated  steam  at  34.8  Ibs.  absolute  weighs  .0836  Ibs. 

The  weight  of  .07187  cu.  ft.,  or  the  cushion  steam,  will  then  equal 
.07187  X  .0836  =  .006  Ibs. 

The  total  weight  of  steam  in  the  cylinder  during  expansion  is  therefore 

.0254  +  .006  =  .0314  Ibs. 

Finally  plot  the  curve  of  saturation  for  .0314  Ibs.  of  steam.  To  do 
this,  take  any  pressure  such  as  80  Ibs.  absolute  and  from  the  steam 
tables  find  the  volume  of  1  Ib.  of  steam  at  that  pressure.  This  equals 
5.47  cu.  ft.  The  volume  of  .0314  Ibs.  would  then  be 

.0314  X  5.47  =  .1718  cu.  ft.  =  297  cu.  in. 


\ 

TESTING  OF  STEAM  ENGINES  177 

Hence  the  ordinates  of  this  point  will  be 

en 
P  =  60  =  L33  in. 

297 
and  v  =  2^7  =  1.43  in. 

This  point  is  then  plotted,  and  others  are  found  and  plotted  in  the 
same  way.  A  curve  drawn  through  these  points  will  be  the  saturation 
curve. 

116.  Determination  of  Steam  Consumption. — When  the  engine 
is  used  with  a  surface  condenser,  the  steam  consumption  may  be 
determined  by  weighing  the  steam  condensed.  It  is  seldom, 
however,  that  this  can  be  done,  and  usually  it  is  necessary  to 
measure  the  amount  of  feed  water  going  to  the  boiler  which 
supplies  steam  to  the  engine  to  be  tested.  When  this  is  done, 
great  care  should  be  taken  to  see  that  all  the  steam  produced 
from  this  feed  water  goes  to  the  engine.  If  all  the  steam  does 
not  go  to  the  engine,  the  amount  going  to  other  purposes  should 
be  measured  and  deducted  from  the  total  feed,  the  difference 
being  the  engine  feed.  Tests  of  this  character  should  be  at  least 
10  hours  in  length,  and  still  better  24  hours,  so  as  to  allow  for 
the  effect  of  varying  conditions  such  as  level  of  water  in  the  boiler. 
The  engine  should  be  credited  with  the  moisture  in  the  steam. 
The  engine  should  be  operated  for  some  time  before  the  test  be- 
gins so  that  the  heat  conditions  may  be  uniform.  During  the 
test  the  engine  should  be  run  as  nearly  as  possible  at  a  uniform 
load.  Indicator  cards  are  usually  taken  every  10  or  15  minutes, 
and  the  average  horse-power  shown  by  the  cards  is  taken  as  the 
average  horse-power  developed  during  the  test.  As  has  already 
been  stated,  to  determine  the  number  of  pounds  of  steam  used 
by  a  steam  engine  per  horse-power  per  hour,  the  water  entering 
the  boiler  is  weighed  and  all  the  water  that  actually  goes  to  the 
engine  is  charged  to  the  engine.  This  weight  of  water  reduced 
to  pounds  per  hour  is  divided  by  the  average  horse-power  devel- 
oped by  the  engine;  the  result  is  the  number  of  pounds  of  steam 
used  by  the  engine  per  horse-power  per  hour.  The  American 
Society  of  Mechanical  Engineers  has  adopted  a  standard  method 
of  testing  steam  engines,  which  will  be  found  in  Volume  XXIV  of 
their  Proceedings. 

The  number  of  pounds  of  steam  used  by  the  various  forms  of 
12 


178 


HEAT  ENGINES 


engines  are  summarized  in  the  following  table.     These  results 
are  very  general  for  the  various  classes  of  engines. 

TABLE  XIX.  STEAM  CONSUMPTION  OF  VARIOUS  CLASSES  OF  ENGINES 

Pounds 

Simple  throttling  engine,  non-condensing 44  to  45 

Simple  automatic  engine,  non-condensing 30  to  35 

Simple  Corliss  engine,  non-condensing 26  to  28 

Simple  automatic  engine,  condensing 22  to  26 

Simple  Corliss  engine,  condensing 22  to  24 

Compound  automatic  engine,  non-condensing 25  to  30 

Compound  automatic  engine,  condensing 18  to  20 

Compound  Corliss  engine,  condensing 14  to  16 

Triple  Corliss  engine,  condensing 12.25  to  13 

Uniflow  engine,  simple  condensing,  superheat 11.25  to  12 

117.  Brake  Horse-power. — All  of  the  economies  given  in 
Table  XIX  are  based  on  the  indicated  horse-power  of  the  engines. 
But  this  does  not  represent  the  actual  useful  work  that  can  be 


FIG.  88.— Prony  brake. 

obtained  from  the  engine,  as  part  of  this  power  must  be  used  in 
overcoming  the  friction  of  the  engine  itself.  The  actual  power  of 
the  engine  delivered  upon  the  fly-wheel  is  usually  measured  by 
a  Prony  brake  or  some  similar  device.  The  horse-power  obtained 
at  the  brake  is  termed  the  "brake,"  or  "effective"  horse-power. 
The  brake  used  to  determine  the  brake  horse-power  usually 
consists  of  an  adjustable  strap  which  encircles  the  rim  of  the 
brake  wheel  which  is  fastened  to  the  crank-shaft  of  the  engine. 
The  brake  wheel  should  be  provided  with  interior  flanges  for 
holding  water  used  for  keeping  the  rim  cooled.  To  the  strap 
encircling  the  brake  wheel  is  rigidly  fastened  an  arm  which  rests 
on  a  platform  scales.  The  friction  of  the  strap  DE,  Fig.  88, 


TESTING  OF  STEAM  ENGINES  179 

tends  to  carry  the  arm  FK  in  the  direction  of  rotation  of  the  wheel. 
The  force  tending  to  depress  the  arm^-K  is  measured  on  the  scales. 
The  net  force  on  the  scales  times  the  distance  AC  is  the  moment 
of  friction,  and  this  multiplied  by  the  angular  velocity  equals  the 
rate  of  doing  useful  work.  The  weight  of  the  lever  on  the  scales 
must  either  be  counterbalanced,  or  else  found  by  suspending  the 
lever  on  a  knife-edge  vertically  over  A  and  noting  the  scale 
reading.  This  weight  plus  the  weight  of  the  standard  C  is  called 
the  tare,  and  is  then  subtracted  from  the  weight  shown  on  the 
scales  to  determine  the  net  weight  due  to  the  force  of  friction. 
The  standard  C  must  be  of  such  a  length  that  when  the  engine 
is  running  the  arm  FK  is  held  in  a  horizontal  position. 

Let  w  =  the  net  weight  on  the  scales,  n  the  revolutions  of 
the  shaft  per  minute,  I  the  horizontal  distance  AC  in  feet,  or  the 
brake  arm,  and  B.H.P.  the  brake  horse-power.  Then 

RHP        2irlwn-  m 

:    33000 

118.  Mechanical  Efficiency. — The  brake  horse-power  divided 
by  the  indicated  horse-power  is  the  mechanical  efficiency  of  the  en- 
gine, and  the  indicated  horse-power  minus  the  brake  horse-power 
is  called  the  friction  horse-power.  The  mechanical  efficiency 
of  an  engine  is  usually  about  85  per  cent.,  and  in  well-built 
engines  may  be  as  high  as  90  per  cent,  and  over. 

In  large  engines  it  is  not  possible  to  obtain  the  brake  horse- 
power, as  such  an  engine  would  require  a  very  elaborate  brake. 
In  such  cases  it  is  customary  to  obtain  the  horse-power  lost  in 
friction,  approximately,  by  what  is  termed  a  friction  card. 
A  friction  card  is  obtained  by  removing  all  the  load  from  the 
engine  so  I/hat  the  only  load  acting  upon  the  engine  is  the  friction 
of  the  engine  itself.  An  indicator  card  is  taken  from  the  engine 
under  these  conditions,  and  the  horse-power  shown  by  this  card 
is  called  the  friction  horse-power.  A  card  so  taken  does  not  give 
the  actual  friction  of  the  engine,  as  the  friction  increases  with  an 
increase  of  load.  After  finding  the  friction  horse-power,  the 
actual  output  of  the  engine  may  be  determined  by  subtracting 
this  friction  horse-power  from  the  indicated  horse-power.  If 
the  power  taken  by  the  friction  card  is  more  than  10  per  cent,  of 
the  full-load  capacity  of  the  engine,  the  friction  of  the  engine  is 
considered  to  be  excessive.  Where  an  engine  is  used  to  drive  a 
dynamo,  the  mechanical  efficiency  of  the  engine  may  be  deter- 


180  HEAT  ENGINES 

mined  from  the  electrical  output  of  the  generator,  if  the  electrical 
efficiency  of  the  generator  is  known. 

119.  Actual  Heat  Efficiency.  —  The  actual  thermal  efficiency  of 
an  engine  is  the  heat  equivalent  of  one  horse-power  per  hour  divided 
by  the  number  of  heat  units  consumed  by  the  engine  per  H.P.-hour, 
either  indicated  or  brake. 

Since  a  horse-power  is  33,000  foot-pounds  per  minute,  then  the 
heat  equivalent  of  one  horse-power  per  hour  is 


Let  S  equal  the  steam  consumption  of  an  engine  per  horse- 
power per  hour,  q  the  quality  of  the  steam,  L  the  latent  heat,  h 
the  heat  of  the  liquid  above  32°,  and  t  the  temperature  of  the 
feed-water.  (The  British  practice  assumes  this  temperature 
to  be  the  temperature  corresponding  to  the  exhaust,  or  back, 
pressure.)  Then  the  actual  thermal  efficiency  would  be 

:  ____  2s45  M 

S[h+qL-  (J-32)}' 

120.  Duty.  —  The  economy  of  pumping  engines  is  usually 
expressed  not  as  the  number  of  pounds  of  steam  per  I.H.P.  per 
hour,  but  in  terms  of  "duty." 

In  the  earlier  history  of  pumping  engines,  the  definition  of  duty 
was  the  number  of  foot-pounds  of  work  done  in  the  pump  cylin- 
der per  100  Ibs.  of  coal  burned  in  the  boiler.  The  objection  to 
this  method  of  determining  duty  is  that  it  includes  both  boiler 
and  engine  economy.  In  purchasing  a  pumping  engine  it  was 
necessary  to  allow  the  contractor  to  furnish  the  boilers  also. 

To  obviate  this  difficulty  it  is  better  to  express  duty  as  the 
number  of  foot-pounds  of  work  obtained  in  the  pump  cylinders 
per  1000  pounds  of  steam  furnished  to  the  engine.  The  speci- 
fications state  at  what  pressure  this  steam  must  be  furnished. 

Duty  may  also  be  expressed  as  the  number  of  foot-p'ounds  of 
work  done  in  the  pump  cylinders  per  1,000,000  B.T.U.  consumed 
by  the  engine.  This  is  the  best  way  of  expressing  duty,  as  it 
eliminates  all  considerations  of  the  steam  pressure.  Engines 
working  under  widely  different  conditions  may  be  compared 
when  their  duty  is  based  on  foot-pounds  developed  in  the  pump 
cylinder  per  1,000,000  B.T.U.  furnished  to  the  engine. 

The  amount  of  "work  done"  is  equal  to  the  weight  of  water 
pumped  times  the  "head"  pumped  against.  The  total  head  is 
made  up  of  the  pressure  shown  by  the  gage  on  the  discharge 
line  plus  that  on  the  suction  fine,  both  reduced  to  feet,  plus  the 


TESTING  OF  STEAM  ENGINES 


181 


vertical  distance  between  the  center  of  the  pressure  gage  and 
the  point  of  attachment  of  the  suction  gage  to  the  main. 

The  duty  that  may  be  obtained  in  the  various  forms  of  pumping 
engines  is  given  in  the  following  table: 

TABLE  XX.  DUTY  OF  VARIOUS  FORMS  OF  PUMPS 

Ft.  Ibs. 

Small  duplex  non-condensing  pumps 10,000,000 

Large  duplex  non-condensing  pumps 25,000,000 

Small  simple  fly-wheel  pumps,  condensing 50,000,000 

Large  simple  fly-wheel  pumps,  condensing 65,000,000 

Small  compound  fly-wheel  pumps,  condensing 85,000,000 

Large  compound  fly-wheel  pumps,  condensing 120,000,000 

Large  triple-expansion  fly-wheel  pumps,  condensing 165,000,000 

Large  triple-expansion     pumps,     condensing,     of     exceptional 

economy 180,000,000 

The  capacity  of  a  pump  is  the  number  of  gallons  pumped  in 
24  hours. 


100 

Iw 

3   80 

P< 

W  70 


a  30 

I 

M  20 
10 


10       20       30 


50       60       70       80       90      100      110     120     130 
Indicated  Horse  Power 
FIG.  89. — Curves  showing  steam  consumption. 

121.  Variation  of  Steam  Consumption.— Most  engines  work 
at  a  varying  load,  so  that  it  is  important  to  know  the  steam  con- 
sumption of  the  engine  at  the  different  loads.  Fig.  89  shows  the 
variation  of  steam  consumption  in  a  100  horse-power  engine  at 


182  HEAT  ENGINES 

various  loads.  The  upper  curve  shows  the  steam  consumption 
when  the  engine  was  running  non-condensing,  and  the  lower  curve 
when  it  was  running  condensing. 

In  these  curves  the  ordinates  represent  the  steam  consumption 
per  horse-power  per  hour,  and  the  abscissae  represent  the  indi- 
cated horse-power. 

Example. — The  area  of  the  indicator  card  from  the  head  end  of  an 
8"  X  12"  double-acting  steam  engine  running  227  r.p.m.  is  1.34  sq.  in., 
and  from  the  crank  end  1.16  sq.  in.  The  length  of  both  cards  is  2.19 
in.,  and  the  scale  of  the  spring  used  was  60  Ibs.  The  diameter  of  the 
piston  rod  is  1|  in.  A  Prony  brake  was  attached  to  the  engine  and  the 
gross  weight  on  it  was  103.5  Ibs.  The  length  of  the  brake  arm  is  54 
in.,  and  the  tare  28.5  Ibs. 

Find  the  (a)  I.H.P.,  (6)  B.H.P.,  (c)  F.H.P.,  and  (d)  mechanical 
efficiency. 

Solution. — (a)  The  average  height,  or  mean  ordinate,  of  the  card 
is  equal  to  the  area  divided  by  the  length,  and  this  multiplied  by  the 
scale  of  the  spring  used  will  give  the  mean  effective  pressure.  Hence, 

1   34 
Head     end  =  0'       X  60  =  27.7  Ibs. 

M.E.P.I  J;JJ 

Crank   end  =  «        X  60  =  23.95  Ibs. 


Area 


Head     end  =  3.1416  X  4  X  4  =  50.26  sq.  in. 

Crank   end  =  (3.1416  X  4  X  4)  -  (3.1416  X  .75  X  .75) 


=  48.50  sq.  in. 

Y)  /  fl  7? 

The  indicated  horse-power  for  each  end  equals  oonnn* 
Hence, 


„     ,          ,       27.  7  X  1  X  50.  26X227 
Head     end  =  - 


H  P 

/c-1         ,       23.95X1X48.5X227 
Crank   end  =  ^3000 

Total  I.H.P.  =  9.58  +  8.02  =  17.6. 

(6)  Net  weight  on  brake  =  103.5  -  28.5  =  75  Ibs. 

54 
Length  of  brake  arm  =    ~  =  4.5  ft. 


2V*3    1<l1fi  V  4.    ^  V  997  V  7^ 
/\  o .  iTtiu  /\  t .  o  /\  A^i    /\   iu  ^  .  „ 

=  33000  =  33000" 

(c)  F.H.P.  =  I.H.P.  -  B.H.P.  =  17.6  -  14.6  =  3. 

(d)  Mech.  Eff.  =  V^V  =  i^'A  =  -829  =  82.9  per  cent. 

l.Xl.-T.  l/.D 


TESTING  OF  STEAM  ENGINES  183 

Example. — If  the  engine  in  the  preceding  problem  used  35  Ibs.  of 
dry  steam  per  I.H.P.  per  hour  at  100  Ibs.  pressure  and  exhausted  it 
at  atmospheric  pressure,  find  (a)  the  theoretical  maximum  thermal 
efficiency,  and  (6)  the  actual  thermal  efficiency  of  the  engine  (based 
on  I.H.P.),  if  the  temperature  of  the  feed  water  is  212°  F. 

Solution. — (a)  The  theoretical  maximum  thermal  efficiency  is  the 
efficiency  of  the  Carnot  cycle  working  in  the  limits  given.  Hence, 
theoretical  efficiency 

_  Tl  -  T2  _  (337.9  +  460)  -  (212  '+  460) 

Tl  (337.9+460) 

797.9  -672      _  125^9 
797.9          "  79779 
=  .1575  =  15.75  per  cent. 
(6)  Actual  thermal  efficiency 

2545 

"  S{H  -(t-32)} 

2545 2545 

~  35(1188. 6 -(212 -32)}  "35X1008.6 

2545 
=  35300  =  .0721  =  7.21  per  cent. 

Example.*— A  500  H.P.  engine  pumps  18,000,000  gallons  of  water 
in  24  hours  against  a  total  head  of  70  Ibs.  per  square  inch.  The  steam 
consumption  is  15  Ibs.  per  I.H.P.  per  hour.  Steam  pressure,  100  Ibs.; 
feed  temperature,  120°.  (a)  What  is  the  duty  per  1000  pounds  of  steam? 
(6)  What  is  the  duty  per  1,000,000  B'.T.U.? 

Solution. — (a)  Weight  of  water  pumped  in  24  hours 

=  81  X  18,000,000  =  150,000,000  Ibs. 
Head  pumped  against  =  70  X  2.31  =  161.7  ft. 

Work  done  in  24  hours  =  150,000,000  X  161.7 

=  24,255,000,000  ft.-lbs. 

Work  done  per  hour  =  24,255,000,000  •*•  24 

=  1,010,625,000  ft.-lbs. 

Steam  used  per  hour  =  500  X  15  =  7500  Ibs. 

Duty  per  1000  Ibs.  of  steam  =  1,010,625,000  -5-  7.5 

=  134,750,000  ft.-lbs. 
(b)  Net  heat  supplied  to  engine  per  pound  of  steam 

=  1188.6  -  (120  -  32)  =  1100.6  B.T.U. 
Total  heat  furnished  to  engine  by  boiler 

=  1100.6  X  7500  =  8,255,000  B.T.U. 
Duty  per  1,000,000  B.T.U.  =  1,010,625,000  -s-  8.255 

=  122,400,000  ft.-lbs. 

*A  gallon  of  water  weighs  8|  Ibs. 

A  water  pressure  of  1  Ib.  per  square  inch  equals  a  head  of  2.31  feet. 
One  inch  of  mercury  equals  a  pressure  of  .491  Ibs. 


184  HEAT  ENGINES 


PROBLEMS 

1.  An  engine  is  8"  X  12"  and  runs  250  r.p.m.     The  indicator  card  of 
the  head  end  has  an  area  of  2  sq.  in.  and  of  the  crank  end,  2.5  sq.  in.     Length 
of  both  cards,  3  in. ;  spring,  80  Ibs.     Diameter  of  the  piston  rod,  1  \  in.     What 
horse-power  does  the  engine  develop? 

2.  An  8"  X  12"  engine  runs  250  r.p.m.     The  indicator  card  from  the 
head  end  has  an  area  of  1.5  sq.  in.  and  length  of  3  in.;  from  the  crank  end  an 
area  of  1.7  sq.  in.  and  length  of  3  in.     The  scale  of  spring  is  80  Ibs.     Diameter 
of  piston  rod,  2  in.     What  horse-power  is  the  engine  developing? 

3.  A  double-acting  engine  is  12"  X  12"  and  runs  250  r.p.m.     The  area 
of  the  average  indicator  card  is  1.5  sq.  in.  and  the  length  is  3  in.     Scale  of 
spring,  60  Ibs.     What  is  the  I.H.P.  of  the  engine? 

4.  The  area  of  the  indicator  card  on  the  head  end  of  an  engine  is  2.3  sq.  in. ; 
area  of  crank  end  card,  2  sq.  in.;  length  of  each  3  in.     Scale  of  spring,  80  Ibs. 
Engine  is  18"  X  24"  and  runs  100  r.p.m.     Diameter  of  piston  rod,  3  in. 
What  is  the  I.H.P.  of  the  engine? 

6.  The  indicator  card  from  the  head  end  of  an  engine  is  2.1  sq.  in.  in  area 

and  3  in.  long;  from  the  crank  end  the  area  is  2.2  sq.  in.  and  the  length  3  in. 

The  cards  were  taken  with  a  100-lb.  spring.     The  engine  is  18"  X  24"  and 

runs   150  r.p.m.     Piston  rod  is  3  in.  in  diameter.     What  horse-power  is 

Jhe  engine  developing? 

6.  An  engine  is  18"  X  36";  r.p.m.,  100;  diameter  of  piston  rod,  3  in. 
Area  of  head  end  card,  3  sq.  in.;  length,  2.5  in.;  area  of  crank  end  card, 
2.8  sq.  in.;  length,  2.5  in.;  scale  of  spring,  60  Ibs.     Find  the  I.H.P. 

7.  An  engine  is  24"  X  36"  and  runs  100  r.p.m.     The  diameter  of  the 
piston  rod  is  4  in.     The  area  of  the  head  end  card  is  1.5  sq.  in.  and  the  length 
3.2  in.     Area  of  crank  end  card  is  1.7  sq.  in.  and  length  3.5  in.     Scale  of 
spring,  100  Ibs.     Find  the  I.H.P. 

8.  The  indicator  card  from  the  head  end  of  an  engine  is  2.1  sq.  in.  in  area 
and  3  in.  long.     From  the  crank  end  it  is  1.8  sq.  in.  in  area  and  3  in.  long. 
The  card  is  taken  with  an  80-lb.  spring.     The  engine,  24"  X  36",  runs 
100  r.p.m.;  piston  rod,  4  in.  in  diameter.     What  horse-power  is  the  engine 
developing? 

9.  A  12"  X  15"  engine  runs  250  r.p.m.     The  area  of  the  head-end  card 
is  1.314  sq.  in.;  of  the  crank-end  card,  1.168  sq.  in.;  the  length  of  each  being 
2.92  in.     The  cards  are  taken  with  a  50-lb.  spring.     Diameter  of  piston  rod, 
2  in.     The  engine  is  fitted  up  with  a  Prony  brake  with  an  arm  4  ft.  9  in. 
long.     The  tare  of  the  brake  is  25  Ibs.  and  the  gross  weight  on  it,   178  Ib. 
Find  the  I.H.P.;  B.H.P.;  F.H.P.;  and  the  mechanical  efficiency. 

10.  A  12"  X  15"  engine  runs  240  r.p.m.  The  area  of  the  head-end  card 
is  1.341  sq.  in.;  of  the  crank-end  card,  1.49  sq.  in.;  the  length  of  each  being 
2.98  in.  The  cards  are  taken  with  a  50-lb.  spring.  Diameter  of  piston  rod, 
2  in.  The  engine  is  fitted  with  a  Prony  brake  having  an  arm  4  ft.  9  in.  long. 
The  tare  of  the  brake  is  23  Ibs.  and  the  gross  weight  on  it,  213  Ibs.  Find 
the  I.H.P.;  B.H.P.;  F.H.P.;  and  the  mechanical  efficiency. 
p  11.  An  8"  X  12"  engine  runs  220  r.p.m.  Area  of  head-end  card  is 
2.1  sq.  in.;  of  the  crank-end  card,  2.04  sq.  in.;  the  length  of  each  being  2.91  in. 
The  cards  are  taken  with  a  40-lb.  spring.  Diameter  of  piston  rod,  1.5  in. 


TESTING  OF  STEAM  ENGINES  185 

The  engine  is  fitted  with  a  Prony  brake  having  an  arm  54  in.  long.  The  tare 
of  the  brake  is  29.25  Ibs.  and  the  gross  weight  on  it,  100.25  Ibs.  Find  the 
I.H.P.;  B.H.P.;  F.H.P.;  and  the  mechanical  efficiency  of  the  engine. 

12.  An  8"  X  12"  engine  runs  221  r.p.m.     Area  of  head-end  card,  2.32 
sq.  in.;  of  the  crank-end  card,  2.34  sq.  in.;  the  length  of  each  being  2.84  in. 
The  cards  are  taken  with  a  40-lb.  spring.     Diameter  of  piston  rod,  1.5  in. 
The  engine  is  fitted  with  a  Prony  brake  having  an  arm  54  in.  long.     The  tare 
of  the  brake  is  20.25  Ibs.  and  the  gross  weight  on  it,  120,25  Ibs.     Find  the 
I.H.P.;  B.H.P.;  F.H.P.;  and  mechanical  efficiency. 

13.  An  engine  uses  14  Ibs.  of  steam  per  I.H.P.  per  hour.     Initial  steam 
pressure,   125  Ibs.;  feed  temperature,   125°  F.     What  is  the  actual  and 
theoretical  thermal  efficiency  of  the  engine? 

14.  An  engine  uses  25  Ibs.  of  steam  per  I.H.P.  per  hour.     Steam  pressure, 
100  Ibs.;  feed  temperature,  200°  F.     Find  the  actual  and  the  theoretical 
thermal  efficiency  of  the  engine. 

16.  An  engine  uses  30  Ibs.  of  steam  per  I.H.P.  per  hour.  Steam  pressure, 
120  Ibs.;  feed  temperature,  125°  F.  Find  the  actual  and  theoretical  thermal 
efficiency  of  the  engine. 

16.  An  engine  uses  35  Ibs.  of  steam  per  I.H.P.  per  hour.     Initial  steam 
pressure,  100  Ibs.;  feed  temperature,  200°  F.     Find  the  actual  and  theoret- 
ical thermal  efficiency  of  the  engine. 

17.  An  engine  uses  24  Ibs.  of  steam  per  I.H.P.  per  hour.     Steam  pressure, 
100  Ibs.;  feed  temperature,  125°  F.     What  is  the  actual  and  theoretical 
thermal  efficiency? 

18.  Given  a  500  k.w.  generating  set;  efficiency  of  the  engine  and  generator, 
85  per  cent.     Steam  pressure,  150  Ibs.;  feed  temperature,  180°.     The  engine 
uses  20  Ibs.  of  steam  per  I.H.P.  per  hour.     Evaporation  from  and  at  212° 
is  10  Ibs.  of  water  per  pound  of  dry  coal.     Dry  coal  contains  13,000  B.T.U. 
per  pound.     What  is  the  heat  efficiency  of  the  plant? 

19.  A  500  I.H.P.  engine  is  direct  connected  to  a  generator.     Efficiency 
of  engine  and  generator  is  85  per  cent.     Engine  uses  10,000  Ibs.  of  steam  per 
hour.     Steam  pressure,  150  Ibs. ;  feed  temperature,  180°.     Evaporation  from 
and  at  212°  per  pound  of  dry  coal  is  10  Ibs.     Dry  coal  contains  13,000  B.T.U. 
per  pound.     What  is  the  heat  efficiency  of  this  plant? 

20.  A  pumping  engine  pumps  15,000,000  gal.  of  water  per  day  (24  hours) 
against  a  head  of  70  Ibs.  per  square  inch.     It  uses  6000  Ibs.  of  steam  per  hour. 
Steam  pressure,  125  Ibs.;  feed  temperature,  150°  F.     What  is  the  duty  per 
million  B.T.U.? 

21.  A  pumping  engine  pumps  15,000,000  gal.  of  water  in  twenty-four 
hours  against  a  head  of  65  Ibs.  per  square  inch.     It  develops  450  H.P.  with  a 
steam  consumption  of  13  Ibs.  per  I.H.P.  per  hour,     (a)  What'  is  the  duty  per 
1000  Ibs.  of  steam?     (6)    What  is  the  mechanical  efficiency?     (c)    If  the 
steam  pressure  is  125  Ibs.  and   the  feed   temperature   130°,  what   is   the 
duty  per  1,000,000  B.T.U.? 

22.  An  engine  develops  450  I.H.P.  and  uses  6300  Ibs.  of  steam  per  hour. 
It  pumps  600-,000  gallons  of  water  an  hour  against  a  head  of  70  Ibs.     (a) 
What  is  the  mechanical  efficiency  of  the  engine  and  pump?     (6)  What  is  its 
duty  per  1000  Ibs.  of  steam?     (c)  If  the  initial  steam  pressure  is  125  Ibs. 
and  the  feed  temperature  is  130°,  what  is  its  duty  per  million  heat  units? 


186  HEAT  ENGINES 

23.  A  20,000,000  gallon  pumping  engine  pumping  against  a  head  of  70 
Ibs.  has  a  duty  of  120,000,000  foot-pounds.     If  the  steam  pressure  is  180 
Ibs.  and  feed  temperature  180°,  how  many  pounds  of  steam  will  be  used 
per  hour? 

24.  A  40,000,000  gallon  pumping  engine  pumping  against  a  head  of  70 
Ibs.  has  a  duty  of  160,000,000  foot-pounds.     If  the  steam  pressure  is  180 
Ibs.  and  feed  temperature  180°,  what  boiler  horse-power  will  be  required 
for  the  plant? 

26.  A  40,000,000  gallon  pumping  engine  has  a  duty  of  120,000,000  toot- 
pounds  per  1,000,000  heat  units.  Steam  pressure,  150  Ibs.  absolute;  ex- 
haust 2  Ibs.  absolute;  feed  temperature,  120°  F.;  pressure  pumped  against, 
70  Ibs.  per  square  inch  gage.  What  boiler  horse-power  will  be  required  to 
operate  the  pump? 

26.  The  duty  of  a  12,000,000  gallon  pumping  engine  is  160,000,000  foot- 
pounds. Steam  pressure,  180  Ibs.  absolute;  exhaust,  2  Ibs.  absolute;  feed 
temperature,  120°  F.;  reading  of  pressure  gage,  60  Ibs.  per  square  "inch; 
reading  of  suction  gage,  20  in.  of  mercury;  distance  between  center  of 
pressure  gage  and  point  of  attachment  of  suction  gage,  10  feet,  (a) 
What  boiler  horse-power  will  be  required  to  operate  the  plant?  (6)  If  the 
mechanical  efficiency  of  the  pump  is  90  per  cent.,  what  will  be  the  steam 
consumption  per  I.H.P.  per  hour?  (c)  If  the  boiler  efficiency  is  70  per  cent, 
and  the  coal  contains  13,000  B.T.U.  per  pound  and  costs  $3  per  ton,  what  will 
be  the  coal  cost  per  year,  if  the  plant  operates  twenty-four  hours  a  day  for 
three  hundred  and  sixty-five  days  per  year? 


CHAPTER  XI 
VALVE  GEARS 

122.  AN  essential  part  of  every  steam  engine  is  the  valve. 
The  function  of  the  valve  is  to  admit  steam  to  the  cylinder  at 
the  proper  time  in  the  stroke,  and  on  the  return  stroke  to  open 
the  cylinder  to  the  exhaust  and  let  the  steam  escape  either  to 
the  atmosphere  or  to  the  condenser.  The  proper  action  of  the 
engine  depends  very  largely  upon  the  proper  distribution  of 
steam  in  the  cylinder. 

In  a  single-acting  engine,  steam  is  admitted  to  one  side  of  the 
piston  only,  while  in  the  double-acting  engine  it  is  admitted 
alternately,  first  to  one  side  and  then  to  the  other.  Most  steam 
engines  in  common  use  are  double-acting. 

In  the  simpler  forms  of  steam  engines,  only  one  valve  is  used, 
which  is  so  arranged  that  it  admits  steam  to  either  end  of  the 
cylinder  and  also  controls  the  exhaust. 


FIG.  90.— D-slide  valve. 

123.  Plain  D-slide  Valves.— Fig.  90  shows  a  plain  D-slide 
valve,  so  called  from  its  longitudinal  cross-section.  In  the  figure 
shown,  the  space  D  is  filled  with  live  steam  under  pressure,  and 
the  space  C  is  open  to  the  exhaust.  In  the  position  shown, 
steam  has  just  ceased  flowing  from  the  space  D,  through  the  steam 

187 


188 


HEAT  ENGINES 


port  A,  into  the  cylinder.  On  the  other  side  of  the  piston,  steam 
is  exhausting  through  the  steam  port  B  into  the  exhaust  space  C. 
The  valve  is  moving  to  the  left  and  the  piston  to  the  right,  and 
the  point  of  cut-off  has  just  been  reached.  The  steam  will  now 
expand  in  the  cylinder  until  the  valve  has  moved  far  enough  to 
the  left  to  uncover  port  A,  placing  it  in  communication  with  the 
exhaust  port  C,  when  exhaust  will  begin.  Compression  in  the 
right  end  of  the  cylinder  will  begin  when  the  valve  has  moved 
far  enough  to  the  left  to  cover  port  B.  When  it  has  moved  still 
further  to  the  left,  port  B  will  again  be  uncovered  and  steam  will 
be  admitted  to  the  right  end  of  the  cylinder,  driving  the  piston 
toward  the  left.  A  plain  D-slide  valve  will,  therefore,  if  given  a 
proper  reciprocating  motion,  control  the  admission  and  the  ex- 
haust of  the  steam  so  that  the  piston  will  be  given  a  reciprocating 
motion. 

124.  Lap,  Lead,  Angular  Advance,  and  Eccentricity. — Con- 
sider a  valve  such  as  is  shown  in  Fig.  91.     This  valve  is  con- 


FIG.  91. — Simple  valve  without  lap. 


FIG.  92. — Indicator 
card  and  Zeuner  dia- 
gram from  valve  shown 
in  Fig.  91. 


structed  so  that  it  just  covers  the  steam  ports  A  and  B.  If  the 
valve  is  moved  to  the  right,  or  to  the  left,  steam  will  be  admitted 
to  the  cylinder  at  one  end  or  the  other,  and  exhaust  from  the 
opposite  end.  A  valve  constructed  as  shown  will  admit  steam 
to  the  end  of  the  stroke  and  permit  the  exhaust  to  continue 
to  th&  end  of  the  stroke  at  the  opposite  side  of  the  piston. 
There  would  then  be  no  expansion  of  the  steam  on  the  working 
stroke  and  no  compression  of  steam  on  the  exhaust  stroke. 


VALVE  GEARS 


189 


The  ideal  indicator  card  for  a  valve  such  as  is  shown  in  Fig.  91 
is  shown  in  Fig.  92. 

In  certain  steam  pumps,  the  indicator  card  is  very  similar  to 
the  one  shown.  The  economy,  however,  of  such  a  pump  must 
be  very  poor.  In  order  partially  to  expand  the  steam  and  to 
have  compression  at  the  end  of  the  exhaust  stroke,  it  is  neces- 
sary that  the  valve  be  lengthened  as  shown  in  Fig.  93.  The 
lengthening  of  the  valve  on  the  steam  side  causes  the  port  to  be 
closed  before  the  end  of  the  stroke,  and  for  the  balance  of  the 
stroke  the  steam  expands.  The  increased  length  of  the  valve 
on  the  steam  side  of  the  valve  is  called  the  steam  lap.  The 
steam  lap  in  Fig.  93  is  the  distance  S.  Steam  lap  may  be  defined 
as:  the  distance  that  the  valve,  when  in  its  mid-position,  extends 
beyond  the  edge  of  the  steam  port  toward  that  side  from  which  it  takes 
steam.  It  is  equal  to  the  distance  the  valve  must  move  from  its 
mid-position  before  steam  is  admitted  to  the  cylinder.  The  lap 
is  not  always  the  same  for  the  two  ends  of  the  cylinder. 


FIG.  93. — Valve  with  steam  and  exhaust  lap. 

In  order  to  have  compression  at  the  end  of  the  exhaust  stroke, 
the  valve  must  extend  beyond  the  exhaust  port  as  shown  in  Fig. 
93,  by  an  amount  called  the  exhaust  lap,  which  may  be  denned 
as  follows:  exhaust  lap  is  the  distance  that  the  valve,  when  in  its 
mid-position,  extends  beyond  the  edge  of  the  steam  port  toward  that 
side  into  which  it  exhausts.  It  is  equal  to  the  distance  the  valve 
must  move  from  its  mid-position  before  exhaust  begins.  The 
exhaust  lap  may  be  negative,  in  which  case  it  is  equal  to  the  amount 
the  port  P,  Fig.  93,  would  be  open  to  the  exhaust  chamber  C, 
when  the  valve  is  in  the  mid-position. 

If  the  valve  did  not  begin  to  admit  steam  until  just  as  the  engine 
was  on  the  dead  center,  full  steam  pressure  in  the  cylinder  would 
not  be  attained  until  the  piston  had  travelled  some  distance  on 
the  next  stroke.  Jn  order  to  have  full  steam  pressure  in  the 


i 


190  HEAT  ENGINES 

cylinder  at  the  beginning  of  each  stroke,  it  is  necessary  for  the 
valve  to  open  just  before  the  piston  reaches  the  end  of  the  previous 
return  stroke,  thus  causing  pre-admission.  This  opening  before 
the  end  of  the  stroke  is  called  the  lead. 

Lead  is  the  amount  the  steam  port  is  open  when  the  piston  is  at 
the  end  of  its  stroke. 

If  the  valve  were  to  be  constructed  as  shown  in  Fig.  91,  the 
eccentric  would  be  set  exactly  90°  in  advance  of  the  position 
of  the  crank.  But  with  the  valve  having  both  lap  and  lead,  it 
is  necessary  to  set  the  eccentric  ahead  of  the  crank  an  angle 
greater  than  90°  by  an  amount  sufficient  to  "move  the  valve  a 
distance  equal  to  the  lap  and  the  lead.  This  angle  is  called  the 
angle  of  advance. 

The  angle  of  advance  is  the  angle  which  the  perpendicular  to  the 
line  of  motion  of  the  piston  makes  with  the  center  line  of  the  eccentric 
when  the  engine  is  on  the  dead  center;  or  it  is  the  angle  between  the 
center  lines  of  the  eccentric  and  the  crank  minus  90°. 

Eccentricity  is  the  distance  between  the  center  of  the  shaft  and 
the  center  of  the  eccentric. 

The  throw  of  the  eccentric  is  equal  to  the  travel  of  the  valve,  or  to 
twice  the  eccentricity. 

125.  Relative  Position  of  Valve  and  Piston. — In  order  to 
study  the  action  of  the  valve,  it  is  necessary  to  know  its  exact 


FIG.  94. — Relative  position  of  valve  and  piston. 

position  for  each  position  of  the  piston.  The  valve  is  driven  by 
an  eccentric  (see  paragraph  108),  which  is  really  a  crank  in  which 
the  crank-pin  is  enlarged  until  it  includes  the  shaft.  As  the  size 
of  the  crank-pin  has  nothing  to  do  with  the  motion  produced  in 
the  rod  attached  to  it,  any  two  cranks  having  the  same  arm,  or 
distance  between  the  shaft  center  and  the  crank-pin  center,  will 
produce  the  same  motion.  In  the  eccentric  the  arm  is  called 
the  eccentricity.  As  the  eccentric  is  equivalent  to  a  crank,  our 


VALVE  GEARS  191 

problem  consists  in  finding  the  simultaneous  positions  of  two 
reciprocating  pieces,  driven  by  two  cranks  upon  the  same  shaft. 

In  Fig.  94  let  OC  represent  any  position  of  the  crank,  OD  the 
center  line  of  the  eccentric,  a  the  angle  between  the  two,  and  BC 
the  connecting  rod.  Drawing  the  arc  CI  with  B  as  a  center,  we 
find  that  the  piston  has  moved  a  distance  El  from  its  extreme 
position  at  the  left,  or  its  distance  from  its  mid-position  is  01. 
Similarly,  by  dropping  a  perpendicular  from  D  upon  EH,  the 
valve  is  found  to  be  at  the  distance  UG  from  its  extreme  right 
position,  or  distance  OU  from  its  mid-position. 

To  be  absolutely  correct  an  arc  should  be  struck  through  D 
with  a  radius  equal  to  the  length  of  the  eccentric  rod  and  with 
the  center  on  the  line  OB,  or  OB  extended  to  the  left,  and  the 
point  U  found  as  the  intersection  of  this  arc  with  the  line  EH, 
rather  than  as  the  foot  of  the  perpendicular  dropped  from  D. 
However,  as  the  ratio  of  the  length  of  the  eccentric  rod  to 
the  length  of  the  eccentric  arm,  or  eccentricity,  is  so  great,  the 
error  caused  by  using  the  perpendicular  instead  of  the  arc  is 
negligible. 

126.  Valve    Diagrams. — The    above    method,    although    the 
most  apparent  way  of  attacking  the  problem,  is  inconvenient  in 
practice,  and  simpler  constructions  known  as  valve  diagrams 
are  commonly  used.     Of  the  many  forms  of  valve  diagrams 
which  have  been  designed,  the  one  due  to  Zeuner  is  perhaps  the 
best  known,  and  will,  therefore,  be  used  in  the  present  discussion. 

127.  Zeuner  Diagram.— Let  XY,  Fig.  95,  represent  the  stroke 
of  the  piston  and  DE  the  travel  of  the  valve.     Let  OA  represent 
any  position  of  the  crank,  the  cylinder  being  assumed  to  be  to 
the  left  of  the  figure;  OB,  the  corresponding  position  of  the  eccen- 
tric; and  a  the  angle  between  the  crank  and  the  eccentric.     Then 
the  angle  BOJ,  or  d,  is  the  angle  of  advance  (a  —  -90°). 

Drop  a  perpendicular  from  B  to  the  line  DE.  Then  OC  repre- 
sents the  displacement  of  the  valve  from  the  mid-position  when 
the  crank  is  in  the  position  OA.  Draw  OF  so  that  the  angle 
FOG  equals  the  angle  of  advance,  5  (  =  a  -  90°).  Then 

FOA  =      90°  -  DO  A  -  FOG 
=    90°  -  B  -  (a  -  90°) 
=  180°  -  0  -  a. 
But  180°  -  6  -  a  =  BOE. 
Therefore  FOA    =  BOE. 


192 


HEAT  ENGINES 


As  a  is  a  constant,  this  relation  will  hold  for  all  values  of  6. 
Draw  FH  from  F  perpendicular  to  OA.  Then  the  triangles 
FOR  and  BOG  are  equal  and  OH  will  equal  OC,  or  the  displace- 
ment of  the  valve  from  its  mid-position.  Since  FH  is  perpen- 
dicular to  OA,  FHO  will  be  a  right  angle  for  any  value  of  0,  and 
the  locus  of  the  point  H  will  be  a  circle  described  on  OF  as  a  diam- 
eter. Therefore,  as  OA  represents  any  position  of  the  crank, 
the  corresponding  displacement  of  the  valve  from  its  mid-position 
may  be  found  by  measuring  the  length  intercepted  by  the  circle 
FHO.  Thus  with  the  crank  at  KO,  the  intercept  is  zero  and  the 
valve  is  at  its  mid-position ;  at  PO  the  intercept  is  a  maximum  and 
the  valve  is  at  its  extreme  position  toward  the  right;  at  LO  the 
intercept  is  again  zero  and  the  valve  has  returned  to  its  mid- 


FIG.  95. — Elementary  Zeuner  diagram. 

position.  Beyond  this  point  the  crank  does  not  intersect  the 
circle,  and  it  is  necessary  to  draw  a  second  circle  OMN,  from 
which  we  obtain  the  location  of  the  valve  when  on  the  left  side 
of  its  mid-position.  The  circles  FHO  and  OMN  are  known  as 
valve  circles.  It  is  important  to  note  that  in  the  arrangement 
which  we  have  selected  (clockwise  rotation  with  the  cylinder  at 
the  left  of  the  shaft)  the  intercepts  on  the  crank  line  made  by  the 
upper  valve  circle  represent  the  displacements  of  the  valve 
toward  the  right,  while  those  made  by  the  lower  valve  circle 
indicate  displacements  toward  the  left. 

128.  Effect  of  Lap. — In  a  valve  having  lap,  it  is  evident  the 
valve  will  have  to  be  moved  from  its  mid-position  a  distance 
equal  to  the  lap  before  admission  begins,  and  on  returning  to 


VALVE  GEARS 


193 


its  mid-position  will  close  the  steam  port  when  its  distance  from 
that  position  is  equal  to  the  lap.  The  effect  of  the  lap  is  to  close 
off  the  steam  before  the  end  of  the  stroke  and,  with  a  very  large 
lap  and  a  small  port,  the  time  of  admission  of  steam  might  be 
reduced  to  zero. 

The  amount  by  which  the  port  is  open  at  any  instant  is  called 
the  port  opening,  and  for  the  valve  in  Fig.  91  is  equal  to  the  dis- 
placement of  the  valve  from  its  mid-posifoon.  Since  the  addition 
of  lap  makes  it  necessary  to  move  the  valve  a  distance  equal  to 
the  lap  before  any  port  opening  is  obtained,  the  port  opening  is 
found  by  subtracting  the  lap  from  the  displacement  of  the  valve. 
This  is  most  conveniently  done  by  drawing  the  circular  arc  DBW, 
Fig.  96,  with  a  radius  equal  to  the  steam  lap.  Then  for  any 
crank  position,  as  OE,  the  opening  of  the  port  is  BC.  A  further 


FIG.  96. — Zeuner  diagram,  showing  effect  of  steam  lap. 

examination  of  this  diagram  shows  that  admission  begins  with 
the  crank  at  OF,  and  ends  at  OG.  This  arrangement  is  not 
practicable  since  admission  must  occur  before  the  piston  begins 
its  stroke,  by  the  amount  of  the  lead.  This  result  may  be  ob- 
tained by  revolving  the  circle  DCWO  in  a  counter-clockwise 
directionabout  0  as  a  center  until  D,  the  point  of  intersection 
of  the  valve  circle  with  the  steam  lap  circle,  falls  below  the  hori- 
zontal axis  OL  as  shown  in  Fig.  97. 

In  rotating  the  valve  circle  to  its  new  position,  the  eccen- 
tricity remains  the  same,  but  the  angular  position  of  the  eccen- 
tric relative  to  the  crank  has  been  altered.  To  see  what  change 
in  the  actual  engine  corresponds  to  a  rotation  of  the  valve  circle 
center  to  E,  we  have  only  to  remember  that  FOG,  Fig.  97,  equals 

13 


194 


HEAT  ENGINES 


the  angular  advance,  or  a  —  90°.  Consequently,  any  increase 
in  the  angle  FOG  corresponds  to  an  equal  increase  in  the  angle 
between  the  crank  and  the  eccentric.  An  examination  of  the 
figure  shows  that  with  any  given  lap,  an  increase  in  the  angular 
advance  increases  the  lead  and  makes  cut-off  earlier.  Also, 
that  for  a  given  angle  of  advance  an  increase  in  the  steam  lap 
reduces  the  lead  and  causes  cut-off  to  occur  sooner. 

In  Fig.  97  the  valve  is  in  its  mid-position  when  the  crank  is 
in  the  position  OJ.  This  would  therefore  be  the  crank  position 
at  the  time  of  admission  if  there  were  no  steam  lap.  Similarly 


x  s 

FIG.  97. — Zeuner  diagram  showing  effect  of  lap. 

the  crank  position  at  cut-off  would  be  01.  As  there  is  steam  lap, 
the  crank  positions  at  admission  and  cut-off  are  OH  and  OC. 
These  are  found  by  drawing  lines  from  0  through  D  and  W,  the 
points  of  intersection  of  the  valve  circle  and  steam  lap  circle. 
AB  is  the  lead. 

If  the  valve  had  no  exhaust  lap,  the  port  would  begin  to  open 
to  exhaust  when  the  crank  is  in  the  position  01,  reaching  its 
maximum  opening  at  OK  and  finally  closing  again  at  OJ,  when 
the  valve  resumes  its  mid-position.  Beyond  OJ,  the  continued 
movement  of  the  crank  compresses  the  steam  remaining  in  the 


VALVE  GEARS  195 

left-hand  end  of  the  cylinder  until  OH  is  reached,  when  the  ad- 
mission of  live  steam  begins. 

129.  Exhaust  Lap. — The  valve  shown  in  Fig.  93  is  extended 
on  the  exhaust  side  so  as  to  give  it  exhaust  lap.     The  effect  of 
this  is  similar  to  the  steam  lap  and  causes  the  exhaust  valve  to 
close  before  the  end  of  the  exhaust  stroke,  giving  the  engine 
compression  at  the  end  of  the  exhaust  stroke  of  the  engine.     On 
the  valve  diagram  the  exhaust  lap  is  treated  in  the  same  way 
as  the  steam  lap.     In  Fig.  97,  with  a  radius  OR  equal  to  the 
exhaust  lap,  the  lap  circle  RT  is  drawn,  and  through  the  points 
where  this  circle  cuts  the  valve  circle,  as  T  and  R,  are  drawn 
the  lines  01'  and  OJ',  giving  the  position  of  the  crank  at  the 
time  the  port  is  opened  to  exhaust,  called  the  point  of  release, 
and  at  the  time  the  port  is  closed,  called  the  point  of  compression. 

The  indicator  card  resulting  from  the  steam  distribution 
brought  about  by  the  valve  analyzed  in  Fig.  97  is  shown  below 
the  valve  diagram  and  is  obtained  as  follows: 

Take  any  crank  position,  as  for  example,  OC,  the  cut-off 
position.  Draw  the  arc  CV  with  the  length  of  the  connecting 
rod  as  a  radius  and  with  the  center  on  the  line  OL.  Then  LV 
represents  the  distance  which  the  piston  has  moved  from  the 
beginning  of  its  stroke  up  to  cut-off,  and  V  may  be  projected 
downward,  thus  locating  N.  The  other  points  of  the  diagram 
corresponding  to  the  different  events  are  found  in  a  similar  way. 
The  actual  indicator  card  for  the  engine  we  are  considering 
would  probably  have  its  corners  rounded  off,  due  to  wire  drawing. 
In  examining  actual  indicator  cards,  the  point  of  cut-off  is  the 
point  of  contraflecture  of  the  curve. 

130.  Crank  End  Diagram. — Thus  far  we  have  confined  our 
investigations  to  the  head  end  of  the  cylinder,  but  as  it  is  neces- 
sary to  see  what  takes  place  in  both  ends  of  the  cylinder,  we 
must  be  able  to  draw  the  valve  diagrams  for  the  crank  end  also. 
Fig.  93  shows  that  by  moving  the  valve  to  the  left,  steam  will 
be  admitted  to  the  crank  end  of  the  cylinder  as  soon  as  the  valve 
has  been  moved  a  distance  equal  to  the  steam  lap  on  the  right 
end  of  the  valve. 

Fig.  98  shows  the  valve  diagram  for  both  the  head  and  crank 
ends  of  the  engine.  In  this  figure,  LZ  and  Z'L'  represent  both 
the  stroke  of  the  piston  and,  to  a  different  scale,  the  travel  of  the 
valve.  As  the  lower  circle  of  the  crank  diagram  shows  the  dis- 
placement to  the  left,  the  diagram  for  the  admission  stroke  on 


196 


HEAT  ENGINES 


the  crank  must  be  drawn  below  the  line  O'L'.  Similarly  the 
valve  diagram  for  the  exhaust  end  must  be  drawn  above  the 
line  O'L'. 

131.  Effect  of  Connecting  Rod.— In  Fig.  98  the  lead  has 
been  taken  the  same  for  both  the  head  and  the  crank  end.  It  is 
easily  seen  in  the  figure  that  the  cut-off  at  the  two  ends  of  the 
cylinder  is  not  the  same.  In  drawing  the  arcs  of  a  circle,  CQ 
and  C'Q' ',  to  project  the  positions  of  the  piston  at  the  time  of 


Y" 


FIG.  98. — Zeuwer  diagrams  showing  effect  of  connecting  rod. 

cut-off  upon  the  line  LZ  and  L'Z',  it  will  be  noticed  that  the 
points  Q  and  Q'  fall  on  the  opposite  sides  of  the  feet  of  the  per- 
pendiculars from  the  points  C  and  C',  measured  in  the  direction 
in  which  the  piston  is  moving,  due  to  the  angularity  of  the  con- 
necting rod.  This  difference  in  the  angularity  of  the  rod  on  the 
two  sides  of  the  cylinder  makes  the  cut-off  on  the  crank  end 
less  than  on  the  head  end.  In  order  to  correct  this  inequality 
of  the  cut-offs,  it  is  necessary  to  have  a  smaller  lap  on  the  crank 


VALVE  GEARS 


197 


end  than  on  the  head  end.  The  points  of  compression  and  release 
are  also  made  unequal  for  the  two  ends  of  the  cylinder  by  the 
action  of  the  connecting  rod,  but  can  be  approximately  equalized 
by  a  proper  variation  in  the  exhaust  laps. 

132.  Determination  of  Lap,  Lead,  and  Angular  Advance.— 
Aside  from  its  use  as  a  means  of  exhibiting  the  action  of  the  valve 
in  an  existing  engine,  the  Zeuner  valve  diagram  may  also  be  used 
for  the  purposes  of  design.  This  may  best  be  shown  by  an  ex- 
ample. Let  it  be  required  to  determine  the  steam  and  exhaust- 
laps,  lead,  angle  of  advance,  and  point  of  release,  having  given  the 
points  of  admission,  cut-off  and  compression,  and  the  travel  of  the 
valve. 


FIG.  99. — Use  of  Zeuner  diagram  for  design  of  valve. 


The  lines  OD  and  OW ,  Fig.  97,  are  equal,  being  radii  of  the 
same  circle  and  therefore  the  arcs  OD  and  OW  are  equal. 
Hence  the  arc  DG  is  equal  to  the  arc  WG  and  the  line  G'O 
bisects  the  angle  between  the  crank  positions  at  admission  and 
cut-off. 

Draw  the  circle  MINQ,  Fig.  99,  with  a  diameter,  M N,  equal 
to  the  stroke  of  the  piston,  and  the  circle  XSYW  with  a  diameter, 
XYj  equal  to  the  travel  of  the  valve.  Through  0  draw  IQ  per- 
pendicular to  MN.  Lay  off  on  MN  the  distance  NB  equal  to 
the  per  cent,  of  the  stroke  at  which  admission  begins;  MF,  the 
per  cent,  at  which  cut-off  occurs;  and  ND,  the  per  cent,  at  which 
compression  begins.  Erect  perpendiculars  at  B,  F  and  D  cutting 
the  crank  circle  at  A ,  C  and  P.  (The  angularity  of  the  connecting 


198 


HEAT  ENGINES 


rod  has  been  neglected.)  Draw  the  lines  OA,  OC  and  OP  repre- 
senting respectively  the  crank  positions  at  admission,  cut-off 
and  compression.  Bisect  the  angle  AOC  with  the  line  EOT. 
Then  the  angle  EOS  equals  the  angle  of  advance.  Draw  the 
valve  circles  EJO  and  TLO  on  OE  and  OT  as  diameters,  inter- 
secting the  lines  OA  at  H,  OM  at  U,  OC  at  J,  and  OP  at  L.  Then 
OH  equals  OU  equals  the  steam  lap,  and  OL  equals  the  exhaust 
lap.  With  0  as  a  center  and  the  radius  OH  draw  the  steam  lap 
circle  HVJ  intersecting  the  line  OM  at  V.  Then  UV  will  equal 
the  lead.  With  0  as  a  center  and  the  radius  OL  draw  the  exhaust 
lap  circle  intersecting  the  valve  circle  TLO  at  K.  Draw  the 
line  OKR  from  0  through  K.  This  line  represents  the  crank 
position  at  release.  From  R  drop  the  perpendicular  RG  on  to 
the  line  MN  (neglecting  the  angularity  of  the  connecting  rod). 
Then  MG  equals  the  distance  the  piston  has  travelled  at  the  time 
of  release. 

133.  Piston  Valves  and  Other  Balanced  Valves. — The  plain 
D-slide  valve,  while  entirely  satisfactory  under  certain  concli- 


FIG.  100. — Riding  cut-off,  piston  valve. 


tions,  has  a  number  of  inherent  faults  which  preclude  its  use 
in  many  cases.  Prominent  among  these  is  the  amount  of  resist- 
ance to  movement  which  it  offers  when  used  with  high-pressure 
steam.  An  examination  of  Fig.  90  shows  that  the  entire  back 
of  the  valve  is  exposed  to  live  steam,  with  the  result  that  it  is 
pressed  against  its  seat  with  great  force  and  in  consequence  a 
large  frictional  resistance  must  be  overcome  in  moving  it. 


VALVE  GEARS 


199 


By  using  a  piston  valve,  examples  of  which  are  shown  in  Figs. 
100,  101,  and  102,  this  difficulty  is  overcome,  and,  as  it  is  com- 
monly expressed,  the  valve  is  perfectly  "  balanced,"  since  the 
pressure  upon  the  valve  acts  radially  around  its  entire  circum- 
ference. In  the  plain  D-slide  valve,  leakage  of  steam  past  the 


FIG.   101. — Compound  engine  with  piston  Valve. 

valve  is  prevented  by  the  fact  that  it  is  held  tightly  against  its 
seat  by  the  steam  pressure.  In  the  piston  valve  no  such  force 
is  present,  and  in  stationary  engines  it  is  customary  to  rely  upon 
an  accurate  fit  of  the  valve  for  tightness.  This  makes  it  neces- 
sary to  replace  the  valve  when  the  wear  has  made  the  leakage 
excessive.  In  marine  practice,  tightness  is  obtained  by  the  use 


200 


HEAT  ENGINES 


of  spring  rings  similar  to  those  used  on  a  piston.  So  far  as  the 
valve  diagram  is  concerned,  the  piston  valve  is  the  exact  equiva- 
lent of  the  plain  D-valve,  since  we  may  consider  it  formed  by 
rolling  the  flat  working  surface  of  the  plain  D-valve  into  a  cylin- 
drical form.  The  piston  valve  is  used  extensively  in  marine 


FIG.  102. — Simple  engine  with  piston  valve. 

engines,  compound  locomotives,  and  also  in  a  number  of  types  of 
high-speed  stationary  engines. 

The  engine  shown  in  Fig.  101  is  a  vertical  compound  engine 
having  piston  valves  on  both  low  and  high  pressure  cylinders. 
The  sectional  view  of  the  high-pressure  piston  is  shown  more  in 
detail  in  the  upper  right-hand  corner  of  the  figure.  The  piston 


VALVE  GEARS 


201 


valve  on  the  high-pressure  cylinder  is  a  double-ported  valve. 
The  valves  are  driven  by  a  simple  eccentric.  The  governor  in 
this  engine  is  a  shaft  governor  which  controls  the  action  of  the 
high-pressure  valve. 


FIG.  103. — Double  ported  valve  with  cover  plate. 

Fig.  102  shows  a  simple  engine  with  similar  valve. 

The  valve  often  used  in  high-speed  engines  is  the  one  shown 
on  the  left  in  Fig.  103.  To  the  right  in  Fig.  103  is  shown  the 
cover  plate.  This  cover  plate  is  made  a  scraping  fit  when  it  is 


FIG.  104. — Steam  chest  showing  valve  seat. 

placed  over  the  valve.     This  prevents  any  steam  pressure  on 
top  of  the  valve,  making  it  a  balanced  valve. 

Fig.  104  shows  the  valve  seat  in  the  steam  chest.     The  valve 
and  its  cover  plate  are  fitted  to  this  seat.     The  whole  arrange- 


202 


HEAT  ENGINES 


ment  is  shown  in  cross-section  in  Fig.  55.  The  steam  ports  are 
at  the  ends  of  the  steam  chest,  and  the  exhaust  port  between 
them. 

134.  Double-ported  Valves. — It  will  be  noticed  that  all  valve 
diagrams  so  far  discussed  in  this  text  have  been  drawn  for  a  cut- 
off later  than  half  stroke.  It  is  important  to  notice  the  difference 
introduced  in  the  valve  gear  in  changing  from  a  late  cut-off  to 
an  earlier  one.  If  a  valve  diagram  is  constructed  so  that  the 
cut-off  in  the  cylinder  is  at  f  stroke,  the  results  obtained  will  show 
that  so  short  a  cut-off  in  the  D-slide  valve  is  a  practical  impossi- 
bility. The  eccentricity  and  steam  lap  for  J  cut-off  are  entirely 
too  large  for  practical  use,  although  a  J  cut-off  is  not  extraordi- 
narily early,  but  is  the  working  cut-off  used  in  the  majority  of 
high-speed  engines.  It  is  thus  quite  evident  that  a  simple 


FIG.  105. — Modern  high-speed  engine  valve. 

D-valve  is  not  at  all  suitable  for  early  cut-offs  and  some  modifica- 
tion must  be  made  to  obtain  a  satisfactory  form. 

In  Fig.  105  is  shown  a  modern  high-speed  engine  valve  suit- 
able for  early  cut-off.  At  the  right  end  of  the  valve,  admission 
is  begun  and  the  steam  is  entering  the  port  past  the  end  of  the 
valve  in  the  ordinary  manner.  At  the  same  time  a  flow  of  steam 
past  the  upper  corner  is  taking  place.  This  steam  passes  through 
a  port  in  the  valve  and  enters  the  cylinder  with  the  steam  coming 
directly  past  the  lower  corner.  The  advantage  of  this  arrange- 
ment lies  in  the  fact  that  for  any  given  movement  of  the  valve, 
the  port  opening  is  twice  as  large  as  for  the  simple  D-valve, 
since  we  have  two  ports  instead  of  one.  In  other  words,  for  a 
double-ported  valve  with  a  given  cut-off,  port  opening,  and  lead, 


VALVE  GEARS  203 

the  eccentricity  and  the  steam  lap  are  one-half  as  large  as  for  a 
plain  D-slide  valve  giving  the  same  steam  distribution. 

Another  important  feature  of  this  valve  is  the  pressure  plate 
A,  which  extends  over  the  entire  back  of  the  valve,  thus  relieving 
it  from  the  action  of  the  steam  pressure,  and  consequently  re- 
ducing wear  and  friction.  The  pressure  plate  extends  around 
the  side  of  the  valve  and  rests  upon  the  valve  seat.  It  is  held 
in  its  place  by  a  flat  spring  at  the  back  when  there  is  no  steam 
present  in  the  steam  chest.  In  case  of  a  large  quantity  of  water 
being  present  in  the  cylinder  during  compression,  the  spring 
allows  the  pressure  plate  and  the  valve  to  lift  from  its  seat,  thus 
permitting  the  water  to  escape  instead  of  bursting  the  cylinder 
as  it  would  otherwise  do.  This  form  of  valve  may  be  restored 
to  a  tight  condition  when  worn  by  planing  off  the  faces  of  the 
pressure  plate  which  bear  against  the  valve  seat,  thus  reducing 
the  clearance  between  the  valve  and  pressure  plate. 

The  governor  usually  fitted  in  engines  having  this  type  of 
valve  controls  the  speed  by  alternating  the  point  of  cut-off  as  the 
load  changes.  This  variation  in  cut-off  is  effected  by  changing 
the  position  of  the  eccentric  center  in  one  of  three  ways. 

First. — By  revolving  the  eccentric  around  the  shaft,  thus 
keeping  the  eccentricity  constant  while  varying  the  angle  of 
advance. 

Second. — By  moving  the  eccentric  in  a  straight  line  at  right 
angles  to  the  crank,  thus  altering  both  the  eccentricity  and  the 
angle  of  advance,  but  keeping  the  lead  constant. 

Third. — By  moving  the  eccentric  center  in  a  circular  arc,  the 
center  of  which  is  on  the  opposite  side  of  the  arc  from  the  shaft 
center,  and  very  frequently  directly  opposite  the  crank.  In 
this  case  the  angle  of  advance  and  the  eccentricity  are  both 
varied,  but  not  in  the  same  way  as  in  the  second  type. 

Of  these  three  forms,  the  third  is  the  commonest,  largely 
because  it  is  the  most  convenient.  By  drawing  a  series  of  valve 
diagrams  corresponding  to  the  various  positions  of  the  eccentric 
center,  the  changes  produced  in  the  four  events  of  the  steam 
distribution  are  easily  seen.  With  the  eccentric  swung  from  a 
point  opposite  the  crank,  the  diagrams  show  that  as  the  cut-off 
is  shortened,  the  lead  is  reduced,  while  the  points  of  release  and 
compression  are  made  earlier.  For  cut-off  as  early  as  one-fourth 
stroke,  the  points  of  release  and  compression  are  very  much  too 


204 


HEAT  ENGINES 


early  for  low-speed  engines,  but  not  objectionably  so  for  the  high- 
speed engines  in  which  this  valve  gear  is  used. 

135.  Meyer  Riding  Cut-off. — In  an  engine  having  but  one 
valve,  any  change  in  the  position  of  the  valve  affects  all  the  opera- 
tions of  the  valve.  A  change  in  the  angular  advance  not  only 
changes  the  steam  lead,  but  also  the  exhaust  lead  as  well  as  the 


FIG.  106. — Meyer  riding  cut-off  valve. 

cut-off.  In  the  same  way  the  release  and  compression  depend 
upon  each  other,  and  one  cannot  be  changed  without  changing 
the  other.  In  order  to  regulate  the  speed  of  the  engine  by  the 
cut-off,  it  is  desirable  to  have  some  means  of  changing  the  cut-off 
without  changing  any  other  operation  of  the  valve. 

This  may  be  done  by  having  a  separate  valve  controlling  the 
cut-off.     The  Meyer  valve  is  an  example  of  how  this  may  be  done. 


FIG.  107. — Cross-section  of   Buckeye   valve   gear  for   tandem    compound 

engine. 

Such  a  valve  is  shown  in  Fig.  106.  The  main  valve  is  very  similar 
to  a  D-slide  valve,  except  that  the  steam  ports,  A  A',  pass  through 
the  body  of  the  valve  and  the  steam  enters  the  cylinder  through 
these  ports.  The  cut-off  valve  consists  of  two  blocks  that  are 
fastened  "together  by  a  rod  threaded  through  one  with  a  left- 
hand  thread,  and  through  the  other  with  a  right-hand  thread. 


VALVE  GEARS 


205 


By  turning  this  rod,  the  two  blocks  forming  the  cut-off  valve 
may  be  drawn  together  or  forced  apart.  The  two  valves  are 
operated  by  separate  eccentrics  and  are  so  designed  that  when 
admission  begins,  the  cut-off  valve  does  not  obstruct  the  port 
in  the  main  valve.  At  the  point  of  cut-off,  the  riding  or  cut-off 
valve  covers  the  steam  port  in  the  main  valve.  If  the  cut-off 
valve  blocks  are  moved  farther  apart,  and  the  other  operations 
of  the  valve  are  left  the  same,  the  blocks  will  cover  the  port 
earlier  in  the  stroke,  and  the  point  of  cut-off  comes  earlier. 

In  Fig.  107  is  shown  the  riding  cut-off  valve  used  by  the  Buck- 
eye Engine  Company.  This  is  similar  to  the  Meyer  gear  except 
the  blocks  of  the  riding  cut-off  are  rigidly  fastened  to  each  other. 
The  governor  controls  the  cut-off  valve,  and  a  change  in  the  posi- 
tion of  the  governor  changes  the  relative  position  of  the  two 
va,lves,  so  as  to  shorten  or  lengthen  the  cut-off. 

136.  Corliss  Valves  and  Valve  Gear. — The  Corliss  engine, 
invented  by  George  H.  Corliss  in  1849,  and  in  its  more  recent 


FIG.  108. — Corliss  engine — cylinder  and  frame  section. 

forms  varying  only  slightly  from  the  original  engine  of  this  type, 
is  one  of  the  most  commonly  used  forms  of  reciprocating  engines, 
particularly  in  large  sizes,  in  the  United  States  to-day.  They 
give  as  high  an  economy  as  any  form  of  engine  made.  The  dis- 
tinctive features  of  this  engine  are  the  valves  and  the  valve  gear. 
Valves  of  the  form  shown  in  Figs.  108  and  109  are  used  in 
the  Corliss  engine,  each  end  of  the  cylinder  being  provided  with 
separate  admission  and  exhaust  valves.  Instead  of  sliding  upon 
their  seats  with  a  straight  line  motion  like  a  common  slide  valve, 
these  valves  have  an  oscillatory  motion  about  the  common  axis 
of  the  cylindrical  seat  and  valve.  In  horizontal  cylinders  the 
admission,  or  steam  valves,  are  placed  above  with  their  axes  at 
right  angles  to  the  axis  of  the  cylinder,  while  the  exhaust  valves 
are  similarly  placed  below.  All  four  valves  have  spindles  which 


206 


HEAT  ENGINES 


extend  through  stuffing-boxes  to  the  outside  of  the  cylinder, 
where  they  are  rigidly  connected  to  short  cranks  called  valve 
arms.  As  shown  in  Fig.  109,  these  valve  arms  all  derive  their 
motion  from  the  wrist  plate,  which  is  in  turn  oscillated  by  the 
eccentric  rod.  Valve  rods  permanently  connect  the  arms  of  the 
exhaust  valves  to  the  wrist  plate,  but  for  the  steam  valves  a  trip 
gear  is  provided,  which  disengages  the  valve  arm  at  the  point  of 
cut-off  and  allows  the  valve  to  close  with  a  rapid  motion.  This 
sudden  closure  of  the  valve  is  due  to  its  connection  to  the  dash-pot 
piston.  As  the  valve  opens,  the  dash-pot  piston  is  raised,  pro- 
ducing a  partial  vacuum  in  its  cylinder,  so  that  as  soon  as  the 


To  Governor- 


Inlet 
Valve 


FIG.  109. — Corliss  engine,  showing  arrangement  of  valves. 

trip  gear  releases  the  valve  arm  from  its  connection  with  the 
wrist  plate,  atmospheric  pressure  forces  the  dash-pot  piston 
down  and  closes  the  valve. 

In  Figs.  110  and  111  the  trip  gear  for  the  steam  valve  is  shown. 
The  steam  arm  is  keyed  to  the  valve  stem,  and,  as  the  outer  end 
of  the  arm  is  raised  or  lowered,  the  valve  is  turned  on  its  seat. 
The  knock-off  cam  lever  and  the  bell-crank  lever  are  both  free 
to  oscillate  about  the  valve  stem  as  an  axis.  The  valve  rod  con- 
nects one  end  of  the  bell-crank  lever  with  the  wrist  plate,  and  as 
the  wrist  plate  oscillates  back  and  forth,  the  bell-crank  is  given 
a  rocking  motion  about  the  valve  stem.  The  other  arm  of  the 


VALVE  GEARS 


207 


bell-crank  lever  carries  a  steam  hook,  the  inner  leg  of  which  is 
kept  in  close  contact  with  the  knock-off  cam  lever  by  a  spring. 
In  the  position  shown  in  Fig.  110  the  steam  hook  has  engaged  with 
a  block  on  the  outer  end  of  the  steam  arm,  and,  as  the  valve  rod 
is  moved  to  the  left,  the  steam  hook  is  raised  pulling  up  with  it 
the  outer  end  of  the  steam  arm  and  turning  the  valve  on  its  seat, 
opening  it.  When  the  bell-crank  lever  has  been  turned  about 
its  axis  until  the  point  is  reached  where  the  inner  leg  of  the  steam 
hook  strikes  the  knock-off  cam,  the  outer  leg  will  be  forced  to  the 


Governor  Knock- off  Rod 


Knock- off  Cam  Lever 


Double  Arm  or 
Bell-crank  Lever 


FIG.  110. — Line  diagram  of  Corliss  trip  mechanism. 


right  releasing  the  steam  arm  which  will  suddenly  be  pulled 
downward  by  the  dash-pot  rod  which  is  attached  to  it.  This 
sudden  movement  of  the  steam  arm  closes  the  valve  and  gives  a 
sharp  cut-off.  The  governor  controls  the  position  of  the  knock- 
off  cam,  thus  determining  the  point  at  which  the  steam  hook 
releases  the  valve  arm  and  cut-off  takes  place.  A  safety  cam 
is  provided  so  that  in  case  the  governor  belt  breaks,  the  dropping 
of  the  governor  balls  will  rotate  the  safety  cam  in  a  counter- 
clockwise direction,  causing  cut-off  to  occur  so  early  that  the 
engine  will  stop. 


208 


HEAT  ENGINES 


An  analysis  of  the  motion  of  a  properly  designed  Corliss  valve 
reveals  two  important  points: 

First. — That  the  valve  is  moving  at  nearly  its  greatest  velocity 
when  the  edge  of  the  valve  crosses  the  edge  of  the  port. 

Second. — That  during  the  period  when  the  valve  is  closed  its 
motion  is  very  slight. 

The  first  of  these  features  reduces  the  wire-drawing  effect 
and  makes  the  corners  of  the  indicator  card  more  sharply  defined 
than  is  the  case  with  simple  slide  valves.  The  second  reduces 


FIG.  111. — Corliss  trip  mechanism  for  steam  valve. 

the  friction  and  the  wear,  since  the  valve  is  pressed  against  its 
seat  by  the  full  steam  pressure  during  the  large  part  of  the  period 
when  the  port  is  closed.  The  use  of  the  trip  gear  makes  the  cut- 
off independent  of  all  the  other  events,  and  consequently  the  lead 
and  points  of  compression  and  release  remain  the  same  for  all 
loads.  With  the  Corliss  valve  gear  the  combination  of  excellent 
steam  distribution,  slight  leakage  and  wire  drawing,  with  a  mini- 
mum amount  of  clearance,  is  obtained,  resulting  in  a  high  degree 
of  economy. 


VALVE  GEARS 


209 


A  complete  Corliss  engine  direct-connected  to  an  electric 
generator  is  shown  in  Fig.  112.  This  cut  shows  the  rods  pass 
ing  from  the  governor  to  the  valve  motion.  These  engines  are 
always  side-crank  engines,  having  only  one  bearing  on  the  engine 
frame.  The  end  of  the  shaft  away  from  the  crank  is  supported 
by  a  bearing  separate  from  the  engine  frame,  often  called  the 
" outboard"  bearing.  When  this  bearing  is  on  the  right  side, 
looking  from  the  cylinder  toward  the  fly-wheel,  the  engine  is  said 
to  be  "right-hand;"  when  on  the  left  side,  to  be  "left-hand." 


FIG.  112. — Corliss  engine  and  alternating  current  generator. 

137.  Changing  the  Direction  of  Rotation. — In  all  the  pre- 
ceding valve  diagrams,  the  cylinder  has  been  taken  at  the  left  of 
the  shaft  and  rotation  in  a  clockwise  direction  has  been  assumed. 
Horizontal  engines  rotating  in  this  direction,  or  in  other  words 
taking  steam  in  the  head  end  of  the  cylinder  while  the  crank 
passes  through  the  upper  half  of  its  path,  are  said  to  "run  over." 
To  produce  rotation  in  the  opposite  direction,  or  to  make  the 
engine  "run  under,"  it  is  only  necessary  to  lay  off  the  angle  a  in 
the  opposite  direction  from  the  crank.  That  is,  to  set  the  eccen- 
tric at  an  angle  of  90°  +  5  from  the  crank,  measured  in  a  counter- 
clockwise direction.  By  constructing  the  corresponding  valve 
diagram,  all  the  events  will  take  place  at  the  same  percentage 

14 


210 


HEAT  ENGINES 


of  stroke  as  before,  and  nothing  is  changed  except  the  direction 
of  rotation. 

For  many  purposes  engines  are  required  which  can  be  reversed, 
or  made  to  run  in  either  direction,  at  the  will  of  the  operator. 
By  arranging  the  eccentric  so  that  it  could  be  revolved  through 
an  angle  of  180°  -  25,  the  engine  would  be  made  reversible. 
This  arrangement  has  actually  been  used,  though  it  is  now  prac- 
tically obsolete.  Instead  of  this  construction,  mechanisms 
known  as  reversing  gears  are  used,  which  beside  making  the  engine 
reversible,  permit  a  variation  in  the  point  of  cut-off. 

138.  Stephenson  Link  Motion. — In  1842  Robert  Stephenson 
and  Company  applied  to  their  locomotives  a  form  of  reversing 
gear  which  has  received  the  name  of  the  Stephenson  link  motion. 
This  has  been  more  widely  used  than  any  other  type  of  reversing 
gear.  This  gear,  as  shown  in  Fig.  113,  has  as  its  essential  feature 
a  curved  piece,  or  link,  connected  at  its  ends  to  the  rods  of  the 


FIG.  113. — Stephenson  link  motion. 

two  eccentrics.  On  the  end  of  the  valve  stem  is  a  block,  fitted  to 
slide  in  the  link  and  free  to  turn  on  a  pin  carried  by  the  valve 
stem.  By  means  of  a  bell  crank  and  suspension  rods  connecting 
it  to  the  link,  it  is  possible  to  raise  or  lower  the  link,  and  so  cause 
the  valve  to  take  its  motion  from  any  desired  point  along  the  arc 
of  the  link.  One  end  of  the  link  is  connected  to  an  eccentric  for 
the  " go-ahead"  position,  and  the  other  end  of  the  link  to  an 
eccentric  set  for  the  "back-up"  position.  When  the  block  is 
thrown  to  the  end  controlled  by  the  go-ahead  eccentric,  the  valve 
is  moved  so  as  to  drive  the  engine  forward,  and  when  thrown  to 
the  opposite  end,  the  engine  reverses.  As  the  block  is  moved 
nearer  the  middle  of  the  link,  both  eccentrics  affect  the  motion 
of  the  valve,  and  the  cut-off  is  shortened.  When  the  middle  of 
the  link  is  reached,  admission  and  cut-off  are  found  to  occur  at 


VALVE  GEARS  211 

equal  crank  angles  on  either  side  of  the  dead  center  position  and 
the  engine  has  no  motion.  Beyond  the  mid-position,  the  motion 
of  the  engine  is  in  the  opposite  direction. 

In  American  locomotives  a  rocker  arm  is  always  placed  be- 
tween the  link  block  and  the  valve  stem.  This  arrangement 
causes  the  valve  and  the  link  block  to  move  in  opposite  directions. 
For  this  reason  each  of  the  eccentrics  is  placed  at  an  angle  of  180° 
from  the  position  shown  in  Fig.  113.  In  marine  practice  the 
link  block  is  usually  carried  on  the  end  of  the  valve  stem  as 
shown  in  the  figure. 

139.  Radial  Gears. — In  addition  to  the  Stephenson  Link 
Motion,  a  number  of  other  types  of  reversing  gear  are  in  more 
or  less  common  use.  One  class  of  these,  known  as  radial  gears, 
have  either  one  eccentric  and  derive  part  of  their  motion  from 
the  connecting  rod,  or  are  entirely  without  an  eccentric  and  derive 


ifting  Arm 
Center  of  Lift  STiaft 


FIG.  114. — Diagram  of  Walschaert  valve  gear. 

their  entire  motion  from  the  connecting  rod.  The  most  impor- 
tant of  these  is  the  Walschaert  gear,  which  is  now  being  fitted 
to  a  large  number  of  American  locomotives.  A  diagrammatic 
sketch  of  this  gear  is  shown  in  Fig.  114.  On  the  outer  end  of 
the  crank  pin,  a  second  crank  is  carried,  which  is  connected 
with  the  link  in  such  a  way  as  to  cause  it  to  oscillate  about  its 
point  of  support.  The  valve  stem  is  connected  to  the  vertical 
lever  which  derives  its  motion  both  from  the  block,  carried  on 
the  link,  and  from  the  cross-head  of  the  engine.  By  setting  the 
block  at  different  points  along  the  link,  the  cut-off  may  be  varied 
or  the  engine  reversed.  With  the  Walschaert  gear,  the  lead 
remains  the  same  for  all  cut-offs,  instead  of  increasing  when 
cut-off  is  made  earlier,  as  in  the  Stephenson  gear. 


212  HEAT  ENGINES 

Another  type  of  radial  gear  occasionally  met  with  is  the  Joy 
gear,  shown  in  Fig.  115.  In  this  gear  the  valve  motion  is  derived 
from  the  connecting  rod  through  a  linkage.  The  point  S  is 
permanently  fixed.  With  this  gear  the  steam  distribution  is 
almost  exactly  the  same  for  both  ends  of  the  cylinder,  and  the 
lead  is  constant  for  all  cut-offs. 

Another  method  which  may  be  used  for  reversing  engines 
having  a  balanced  slide  valve  is  to  change,  by  means  of  a  three- 
way  cock,  the  steam  ports  into  exhaust  ports  and  the  exhaust 
ports  into  steam  ports. 


Valve  Stem 


Center  line  of  Cylindej 
Wrist  Pin 


FIG.  115. — Joy  radial  gear. 


140.  Setting  the  Valve  by  Measurement. — In  setting  a  valve, 
the  first  step  is  to  place  the  engine  on  dead  center,  that  is,  the 
piston  at  the  extreme  end  of  its  stroke.  To  do  this,  proceed  in 
the  following  way:  Place  the  engine  near  the  center  and  turn 
it  away  from  the  center  about  15°.  Measure  with  a  tram  from 
a  fixed  point  on  the  frame  to  the  fly-wheel  and  mark  the  wheel. 
While  in  the  same  position,  mark  a  line  across  the  cross-head 
and  the  cross-head  guide.  Now  turn  the  engine  past  the  center 
until  the  lines  on  the  cross-head  and  the  cross-head  guide  again 
coincide.  From  the  same  point  on  the  frame,  mark  the  fly-wheel 
again  with  the  tram.  Bisect  the  distance  between  the  tram 
marks,  and  turn  the  fly-wheel  until  this  point  of  bisection  is  just 
the  length  of  the  tram  from  the  fixed  point  on  the  frame.  The 
engine  will  now  be  on  center.  The  opposite  center  can  be  deter- 
mined in  the  same  wav. 


VALVE  GEARS 


213 


The  next  step  is  properly  to  place  the  valve  on  the  valve  stem. 
The  engine  being  on  center,  move  the  eccentric  on  the  shaft 
until  the  valve  has  a  slight  lead.  Measure  this  lead  very  care- 
fully. Now  place  the  engine  on  the  opposite  center,  and  again 
measure  the  lead.  If  the  lead  is  not  the  same,  move  the  valve 
on  the  stem  one-half  of  the  difference.  Then  repeat  the  opera- 
tion until  the  lead  at  both  ends  is  the  same.  The  valve  is  now 
traveling  equally  over  both  steam  ports.  Now  move  the  eccen- 
tric on  the  shaft,  the  engine  being  kept  on  the  center,  until  the 
port  is  just  closed,  and  then  move  it  ahead  to  the  amount  of  the 
lead  desired.  The  lead  is  set  anywhere  from  "line  and  line"  to 
y\  of  an  inch,  depending  upon  the  speed  and  size  of  the  engine. 

141.  Setting  the  Valve  by  the  Indicator. — It  is  difficult  to  set 
the  valve  exactly  by  measurement.  After  the  valve  has  been 
set  by  measurement,  it  is  best  to  check  the  setting  with  the 
indicator. 


Ad  d'       B 

FIG.  116. — Indicator  card  show- 
ing unequal  distribution  of  work  in 
the  two  ends  of  the  cylinder. 


Ad  B 

FIG.  117.— Indicator  card  showing 
effect  of  insufficient  lead. 


When  the  valve  is  not  set  in  the  proper  position  on  the  stem, 
the  steam  admission  at  the  two  end,s  of  the  cylinder  will  not 
be  alike,  and  the  indicator  card  will  appear  as  shown  in  Fig.  116. 
The  objection  to  this  card  is  that  one  end  of  the  cylinder  is  doing 
more  work  than  the  other.  In  single-valve  engines,  this  condi- 
tion may  be  remedied  by  changing  either  the  position  of  the  valve 
on  the  stem,  or  the  length  of  the  valve-stem.  In  the  Corliss 
engine,  it  is  changed  by  varying  the  relative  length  of  the  governor 
rods  to  the  two  admission  valves. 

Fig.  117  shows  an  indicator  card  taken  on  an  engine  where  the 
valve  has  insufficient  lead.  This  card  can  usually  be  corrected 
by  changing  the  position  of  the  eccentric  on  the  shaft.  The 
eccentric  should  be  changed  in  position  until  the  line  ea  is  a 
vertical  line. 


214 


HEAT  ENGINES 


Fig.   118  shows  an  indicator  card  with  too  much  lead.     As 
before,  this  card  may  be  corrected  by  changing  the  eccentric. 
Fig.  119  shows  an  indicator  card  with  too  much  compression. 


Ad  B      A  d  B 

FIG.    118. — Indicator    card  showing    FIG.    119. — Indicator    card  showing 
effect  of  too  much  lead.  effect  of  too  much  compression. 

In  single-valve  engines  with  automatic  governors  this  often 
occurs  at  light  load.  In  Corliss  engines  it  may  be  corrected  by 
changing  the  length  of  the  rod  connecting  the  valve  and  wrist 
plate. 


FIG.    120. — Indicator   card  showing     FIG.    121. — Indicator  card  showing 


effect  of  "wire-drawing 


effect  of  insufficient  exhaust  lead. 


Fig.  120  shows  an  indicator  card  in  which  the  admission  line 
ab  is  a  falling  line.  This  is  due  to  friction  in  the  admission  valve, 
which  is  usually  caused  by  the  valves  opening  slowly.  With 


FIG.  122. — Indicator  card  showing  effect  of  too  short  cut-off. 

rapidly  opening  valves,  such  as  a  Corliss  valve,  the  admission 
line  will  have  the  dotted  position. 


VALVE  GEARS  215 

The  indicator  card  shown  in  Fig.  121  has  insufficient  exhaust 
lead;  that  is,  the  point  of  release  is  too  late.  With  a  single- 
valve  engine  this  condition  of  exhaust  lead  will  usually  be  ac- 
companied by  insufficient  steam  lead,  and  the  admission  line  will 
be  as  shown  in  the  dotted  position.  Correcting  the  steam  lead 
will  correct  the  exhaust  lead.  In  a  four-valve  engine,  the  lead  of 
the  exhaust  valve  should  be  increased. 

When  an  engine  is  operated  with  a  very  light  load,  the  cut- 
off may  be  so  short  that  the  steam  will  be  expanded  below 
atmospheric  pressure  before  the  valve  opens  to  exhaust.  As 
shown  in  Fig.  122,  this  gives  a  loop  of  negative  work  from  c  to  d, 
and  shows  an  uneconomical  condition  of  operation.  When  this 
occurs  regularly  the  engine  is  too  large  for  the  work  it  has  to  do. 
The  best  way  to  correct  it  is  by  reducing  the  steam  pressure 
until  the  cut-off  is  long  enough  so  that  expansion  is  not  carried 
below  atmospheric  pressure. 


CHAPTER  XII 
GOVERNORS 

142.  In  stationary  engine  practice  it  is  essential  that  the  engine 
operate  at  a  uniform  speed  irrespective  of  the  power  which  it 
develops.     In  most  cases  the  load  on  the  engine  is  continually 
varying,  requiring  a  constant  change  in  the.  amount  of  power 
g^ven  by  the  engine.     There  are  two  general  forms  of  governors 
used  for  this  purpose:  the  throttling  governor  which  regulates  the 
pressure  of  steam  entering  the  engine;  and  the  automatic  or  cut-off 
governor  which  regulates  the  volume  of  steam  admitted,  but  does 
not  change  the  pressure  of  the  steam  entering. 

In  addition  to  the  changes  of  speed  brought  about  by  the 
change  of  external  load  on  the  engine,  there  is  also  a  change  of 
speed  during  each  revolution  of  the  engine  due  to  the  variable 
effort  of  the  steam  on  the  crank  pin  of  the  engine,  and  to  the 
effect  of  the  reciprocating  parts  of  the  engine.  This  variation 
of  speed  is  taken  care  of  by  the  fly-wheel  of  the  engine. 

143.  Throttling  Governors. — In  a  throttling  governor  a  valve, 
usually  of  the  poppet  type  or  other  form  of  balance  valve,  is 
located  in  the  steam  pipe  near  the  engine.     This  valve  is  con- 
trolled by  the  governor  in  such  a  manner  that,  when  the  speed 
of  the  engine  increases,  the  area  of  opening  through  the  valve 
is  reduced,  thereby  increasing  the  velocity  of  the  steam  through 
the  valve  and  reducing  the  pressure  of  steam  entering  the  engine. 
This  governor  regulates  the  speed  of  the  engine  by  varying  the 
pressure  of  the  entering  steam,  the  cut-off  remaining  constant. 

144.  Automatic  or  Variable  Cut-off  Governors. — These  gover- 
nors are  attached  to  the  valve  mechanism  of  the  engine  and,  as 
the  load  on  the  engine  is  reduced,  the  length  of  time  during 
which  steam  is  admitted  to  the  engine  is  reduced  by  making 
the  cut-off  come  earlier.     Thus,  as  the  load  becomes  less,  less 
steam  is  admitted  to  the  engine,  but  the  pressure  of  the  steam 
remains  unchanged. 

145.  Relative  Economy. — The  indicator  cards  shown  in  Fig. 
123  are  taken  from  an  engine  using  a  throttling  governor.     This 

216 


GOVERNORS 


217 


figure  shows  a  number  of  cards  taken  at  different  loads.  Under 
a  light  load,  owing  to  the  action  of  the  governor,  the  steam 
pressure  is  very  low,  while  under  a  heavy  load  the  card  shows 
high  pressure.  At  the  light  load  the  steam  is  expanded  almost 
to  atmospheric  pressure,  but  at  the  heavy  load,  the  cut-off  being 
kept  the  same,  there  is  a  very  small  expansion.  This  condition 
is  not  favorable  to  economical  operation. 

Fig.  124  shows  a  card  similar  to  Fig.  123,  but  taken  from  an 
automatic  engine.  In  this  form  of  governing  the  initial  pressure 
remains  the  same  for  all  loads  and  the  cut-off  varies.  This 
enables  the  engineer  to  select  a  load  giving  a  ciit-off  at  which  an 
engine  using  a  given  steam  pressure  will  show  maximum  economy. 
In  most  engines  this  is  found  to  be  about  one-fourth  stroke; 
therefore  an  automatic  engine  should  be  operated  with  a  load 


x  o 

FIG.  123. — Indicator  card  showing 
effect  of  throttling  governor  when 
load  on  engine  is  varied. 


FIG.  124. — Indicator  card  showing 
effect  of  automatic  governor  when 
load  on  engine  is  varied. 


requiring  the  governor  to  maintain  the  cut-off  as  nearly  as  possi- 
ble at  this  point. 

Under  most  conditions  this  form  of  governor  is  more  econom- 
ical in  its  operation  than  the  throttling  governor.  Actual  ex- 
periment with  an  engine  having  both  an  automatic  and  throt- 
tling governor  shows  the  automatic  governor  to  give  a  steam 
consumption  of  about  75  per  cent,  of  the  steam  consumption 
of  the  same  engine  operated  with  a  throttling  governor. 

146.  Governor  Mechanism. — The  mechanism  of  the  governor 
which  is  to  maintain  the  speed  of  the  engine  uniform  must  be 
such  that  the  change  of  speed  will  cause  a  change  in  the  position 
of  the  parts  of  the  governor.  There  are  two  general  types  of 
mechanism  used  for  this  purpose.  The  fly-ball  governor  is  the 
first  type  and  consists  of  two  balls  fastened  to  pivoted  arms  and 
rotated  by  the  engine,  and  as  the  speed  of  the  engine  increases, 


218 


HEAT  ENGINES 


the  balls  move  out  and  change  either  the  throttle  valve  or  the 
valve  mechanism. 

In  the  second  type,  the  shaft  governor,  the  governor  is  fastened 
to  the  fly-wheel  of  the  engine.  It  usually  consists  of  two  weights 
attached  to  the  fly-wheel  of  the  engine  by  arms.  These  arms 
being  pivoted,  as  the  engine  speed  increases,  the  governor  weights 
move  out  against  the  resistance  of  a  spring.  The  governor  arms 
are  attached  to  the  eccentric,  and  as  the  weights  move  out  the 
position  of  the  valve  changes. 

147.  Fly-ball  Governors. — Fig.  125  shows  a  line  diagram  of 
a  fly-ball  governor.  BB  are  the  balls  of  the  governor.  These 
balls  are  suspended  by  arms  AB,  and  are  also  attached  to  the 


FIG.  125. — Line  diagram  of  a  fly-ball  governor. 


weight  W  by  the  arms  BC.  The  arms  and  balls  of  the  governor 
rotate  around  the  vertical  spindle  AC,  and  are  pivoted  at  the 
point  A.  The  weight  W  is  free  to  move  in  a  vertical  direction 
along  the  axis  AC.  As  the  speed  of  the  engine  increases,  the 
balls  of  the  governor  move  out  into  the  dotted  positions  B'B' '. 
Let  the  force  acting  on  each  one  of  the  balls  in  a  vertical  direc- 
tion be  Pj  w  the  weight  of  each  ball,  and  the  height  through 
which  the  balls  are  lifted,  dh.  The  heavy  weight  W  will  move 
through  a  greater  distance,  kdh.  As  the  work  put  in  must  equal 
the  work  done,  we  have  the  following  equation 


2Pdh  =  2wdh  +  Wkdh. 

k  W 
Therefore  P  =  w  +  —-. 


(1) 


GOVERNORS  219 

(If  the  upper  and  lower  arms  are  the  same  length,  then  k  =  2.) 
The  horizontal  work  will  be  zero.  Let  F  be  the  centrifugal 
force  acting  on  each  ball  to  maintain  it  in  the  dotted  positions; 
then  taking  moments  about  A, 

Pr  =  Fh.  (2) 

Substituting  for  P,  in  equation  (2),  its  value  in  equation  (1), 

and  for  F  the  expression  for  the  centrifugal  force,  —  ,  the  equa- 
tion becomes 

kW 


,  (3) 

where  V  is  velocity  in  feet  per  second. 

0 

wV2 

If  W  =  0,  thenw  =      -h,  or 
gr 

V2        r2 

7  -  i  •  <*> 

This  equation  shows  that  theoretically  the  action  of  the  gover- 
nor is  independent  of  the  weight  of  the  balls.  Practically,  there 
is  considerable  friction  in  the  mechanism  of  the  governor,  and 
the  balls  must  have  considerable  weight  in  order  easily  to  over- 
come the  friction  of  the  governor.  If  the  number  of  revolutions' 
of  the  governor  balls  be  n  per  minute,  then 

«> 

Substituting  in  equation  (4),  the  value  of  F2  as  found  from  (5), 
and  solving  for  n 


n  =  svs  • 

Substituting  equation  (5)  for  F2  in  equation  (3),  and  letting  k  = 
2,  then 


W 

'    w 


2936  (l  +.— )  -  n*h.  (8) 


This  expression  gives  the  relation  of  the  principal  items  of  the 
governor  design. 


220 


HEAT  ENGINES 


148.  Shaft  Governor. — There  are  two  forces  that  may  be  util- 
ized to  control  the  speed  of  an  engine  by  means  of  a  shaft  gover- 
nor. In  the  earlier  form  of  governors,  the  principal  force  was 
centrifugal  force. 


FIG.  126. — Elementary  centrifugal  governor. 


FIG.  127. — Actual  construction  of  centrifugal  governor. 

In  Fig.  126,  the  governor  weight  is  so  suspended  that  it  moves 
approximately  in  a  radial  direction  due  to  the  action  of  centrifugal 
force.  In  the  actual  construction  of  the  governor,  the  centrifugal 


GOVERNORS  221 

force  acts  against  the  resistance  of  a  spring.     In  this  figure,  as 
the  speed  of  the  wheel  increases,  the  centrifugal  force  increases 


FIG.   128. — Elementary  inertia  governor. 

and  the  weight  M  will  move  out  against  the  resistance  of  the 
spring. 

Fig.  127  shows  the  actual  construction  of  a  governor  which  is 


FIG.  129. — Actual  inertia  governor. 

actuated  by  centrifugal  force.     The  governor  in  this  case  regu- 
lates the  position  of  the  eccentric,  as  is  shown  by  the  dotted  lines. 


222  HEAT  ENGINES 

The  angular  advance  and  eccentricity  are  changed  at  the  same 
time,  leaving  the  lead  almost  constant  for  all  positions  of  the 
governor. 

In  Fig.  128  the  weight  M  is  fastened  so  that  centrifugal  force 
has  no  effect  upon  the  movement  of  the  weight,  but  only  produces 
a  stress  in  the  arm  SM.  But,  if  the  wheel  were  suddenly  stopped, 
the  weight  would  continue  to  move,  due  to  the  inertia,  and  exert  a 
force  upon  a  spring  (not  shown)  against  the  resistance  of  which  the 
governor  ball  acts.  The  motion  of  this  weight  is  arranged  to 
change  the  position  of  the  valve.  Inertia  alone  is  not  used  as  the 
actuating  force,  but  a  combination  of  centrifugal  force  and  inertia 
is  used.  Fig.  129  shows  a  form  of  governor  combining  these  two 
forces.  The  two  governor  weights  are  fastened  to  a  single  arm 
which  rotates  around  a  pin  (shown  shaded).  One  weight  has  a 
longer  arm  than  the  other,  arid  is  the  dominating  weight.  As  the 
engine  revolves,  this  weight  tends  to  take  a  radial  position.  This 
action  gives  the  governor  its  initial  position  and  determines  the 
position  of  the  valve.  The  governor  weights  are  suspended  so 
that  if  the  speed  of  the  engine  changes,  the  inertia  of  the  weights 
moves  the  governor  against  one  or  the  other  of  the  stops  shown. 
The  governor  weights  act  against  the  resistance  of  a  spring.  The 
speed  at  which  the  engine  is  to  run  may  be  changed  by  changing 
the  tension  of  this  spring.  The  valve  is  driven  by  a  pin  fastened 
to  the  governor  arm. 

149.  Isochronism. — For  a  given  governor,  w  and  W  are  fixed 
quantities,  and  if  the  governor  is  so  constructed  that  h  is  constant, 
then  n  must  be  constant,  and  the  governor  becomes  isochronous. 
An  isochronous  governor  is  one  in  which  the  balls  are  in  equilib- 
rium at  one  speed  and  only  at  one,  except  for  friction,  and  any 
variation  from  this  speed  will  send  them  to  the  limit  of  their 
travel  in  one  direction  or  the  other.     The  friction  of  the  governor 
makes  it  impossible  for  a  governor  to  be  perfectly  isochronous. 
This  result  is  approximately  obtained  by  using  crossed  arms  so 
that  the  governor  balls  have  a  parabolic  path,  and  the  height  h 
will  remain  approximately  constant.     In  some  forms  of  governors 
the  balls  are  guided  in  a  parabolic  guide  so  that  their  motion  is  an 
exact  parabola  and  give  h  a  uniform  value. 

150.  Hunting. — Over-sensitive    governors    often    exhibit   the 
phenomena  known  as  "hunting."     No  matter  how  quickly  a 
governor  may  change  its  position  in  response  to  a  demand  for 
more  or  less  steam,  the  engine  does  not  respond  instantly.     This 


GOVERNORS  223 

is  in  consequence  of  the  energy  stored  in  the  moving  parts  of  the 
engine,  and  in  the  element  of  time  that  must  elapse  between  the 
moment  when  the  steam  is  admitted  by  the  governor  and  the 
time  that  it  acts  on  the  piston.  Therefore  when  a  sudden  demand 
for  power  is  made  on  an  engine  in  which  the  governor  is  too  sensi- 
tive, or  too  nearly  isochronous,  the  drop  in  speed  will  be  sufficient 
to  force  the  governor  into  a  position  of  over-control,  so  that  too 
much  steam  is  admitted.  This  causes  the  revolutions  tojncrease 
beyond  the  desired  point  and  the  same  over-control  is  exercised  in 
the  opposite  direction.  In  other  words,  the  governor  balls  fly  first 
in  one  direction  and  then  the  other,  "  hunting"  for  the  position  of 
equilibrium:  The  effect  is  to  make  the  speed  of  engine  change 
rapidly,  first  having  an  excess  of  speed,  and  then  a  speed  below  the 
normal.  This  trouble  may  be  overcome  by  adding  a  small  weight 
to  one  of  the  governor  balls,  and  changing  the  tension  of  the 
governor  spring. 

151.  Practical   Considerations. — When   a   properly    designed 
engine  does  not  govern  properly,  the  trouble  is  often  due  to  undue 
friction  in  the  vajve  mechanism,  which  may  be  caused  by  a  tight- 
ening of  the  glands  or  the  journals,  or  by  friction  in  the  dash  pot 
and  springs.     It  may  also  be  due  to  excessive  leakage  in  the  valve, 
unbalancing  it,  or  by  the  valve  being  too  tight.     The  governor 
should  also  be  examined  to  see  that  the  weights  have  not  been 
changed.     The  tension  of  the  springs  should  be  uniform,  if  more 
than  one  spring  is  used. 

If  the  engine  operates  at  a  lower  speed  than  that  desired,  the 
tension  of  the  governor  spring  should  be  increased.  If  this  ten- 
sion has  been  increased  to  the  limit  of  the  spring,  then  additional 
weight  should  be  placed  in  the  governor  balls. 

In  all  forms  of  governors  it  is  necessary  that  the  friction  of  the 
valve  mechanism  be  made  as  small  as  possible,  and  it  should,  if 
possible,  be  a  constant  quantity.  It  is  better  to  have  balanced 
valves,  where  they  are  directly  operated  by  the  governor,  and 
the  valves  should  have  a  small  travel.  In  the  D-slide  type  of 
valve,  small  travel  is  obtained  by  using  a  double-ported  valve. 

In  direct  connected  engines,  2  per  cent,  variation  in  speed  is 
the  maximum  allowable,  and  most  specifications  require  the  va- 
riation to  be  less  than  1  per  cent.  In  mill  engines  a  variation 
of  5  per  cent,  is  sometimes  allowed. 

152.  Fly-wheel. — The  governor  of  an  engine  controls  the  speed 
within  certain  limits  by  controlling  the  action  of  the  valve.     It 


224  HEAT  ENGINES 

takes  a  few  revolutions,  however,  to  bring  the  governor  into 
action. 

The  steam  engine,  however,  has  fluctuations  of  speed  that  occur 
in  the  fraction  of  a  revolution,  and  these  fluctuations  must  be 
controlled  by  the  fly-wheel.  These  fluctuations  of  speed  are  due 
to  three  principal  causes: 

First. — The  pressure  of  steam  is  not  the  same  at  all  points  of  the 
stroke. 

Second. — The  motion  of  the  piston  is  carried  to  the  shaft  by 
means  of  the  connecting  rod  and  crank.  This  means  of  changing 
reciprocating  into  rotary  motion  causes  a  turning  effort  which 
varies  from  zero  to  a  maximum. 

Third. — The  reciprocating  motion  of  the  engine  piston  and 
other  parts  necessitates  these  parts  being  brought  to  rest  and 
started  again  twice  each  revolution.  The  overcoming  of  the 
inertia  effect,  caused  by  the  action  described,  causes  a  variable 
force  to  be  transmitted  to  the  crank. 

A  fly-wheel  is  fastened  to  the  main  shaft  of  the  engine  to  reduce 
the  variation  of  speed  of  the  engine  in  the  fraction  of  a  revolution. 
The  inertia  of  the  fly-wheel  serves  to  carry  the  engine  at  those 
portions  of  the  stroke  where  the  piston  is  not  giving  sufficient 
power  to  the  shaft  to  carry  the  load. 

The  effectiveness  of  the  fly-wheel  depends  upon  the  energy 
stored  in  it.  As  most  of  the  weight  of  the  wheel  is  in  the  rim,  we 
may  consider,  for  an  approximation,  the  action  of  the  rim  as  giving 
the  fly-wheel  effect.  If  W  is  the  weight  of  the  fly-wheel  rim  in 
pounds,  and  R  is  the  average  radius  in  feet,  and  the  wheel  makes 
n  revolutions  per  minute,  then  the  energy  of  the  rim 


W 

2g  V   60 


4rr2 

2 
ft.lb,  (9) 


The  expression  shows  that  the  effectiveness  of  a  fly-wheel  de- 
pends upon  the  weight  of  the  rim,  the  square  of  the  radius  of 
the  wheel,  and  the  square  of  the  number  of  revolutions  that  it 
makes. 


CHAPTER  XIII 
COMPOUND  ENGINES 

153.  Compound  Engines. — Any  engine  in  which  the  expansion 
of  steam  is  begun  in  one  cylinder  and  continued  in  another  is 
called  a  compound  engine]  although  this  term  as  commonly  used 
refers  to  an  engine  in  which  the  expansion  takes  place  in  two 
cylinders  successively.  A  triple-expansion  engine  is  one  in  which 
the  steam  is  expanded  successively  in  three  cylinders. 

When  steam  is  expanded  in  two  or  more  cylinders  successively, 
the  number  of  expansions  per  cylinder  is  less  than  when  only  one 
is  used,  and  therefore  the  range  of  temperature  in  each  cylinder  is 
less.  Reducing  the  range  of  temperature  in  the  cylinder  reduces 
the  condensation  losses.  -The  principal  object  of  compounding  is 
to  reduce  the  amount  of  steam  used  per  horse-power  per  hour,  and, 
under  proper  conditions,  compounding  accomplishes  this,  owing 
to  the  reduction  of  initial  condensation.  The  radiation  losses 
from  a  compound*  engine  are  usually  larger  than  from  a  simple 
engine,  and  very  often  the  mechanical  losses  are  increased  by 
compounding. 

The  tendency,  then,  in  a  compound*  engine,  is  to  increase  the 
radiation  loss  and  to  increase  the  mechanical  losses.  On  the  other 
hand,  compounding  decreases  the  thermodynamic  losses  by 
decreasing  the  range  of  temperature  in  each  cylinder.  With  low 
pressure  and  a  small  number  of  expansions,  a  single-cylinder 
engine  is  more  economical  than  a  compound*  engine,  but  with 
high-pressure  steam  and  a  larger  number  of  expansions,  the 
reverse  is  the  case.  The  higher  the  pressure  and  the  larger  the 
number  of  expansions  the  greater  the  economy  of  the  compound* 
engine. 

For  pressures  under  100  Ibs.,  the  single-cylinder  condensing 
engine  is  more  economical  than  the  compound  engine.  But  for 
pressures  above  100  Ibs.  the  compound  engine  is  usually  more 
economical.  In  the  case  of  the  non-condensing  engine,  the  com- 

*The  term  "compound"  as  here  used  includes  triple-expansion,  quad- 
ruple-expansion, etc. 

15  225 


226 


HEAT  ENGINES 


pound  engine  does  not  show  any  economical  advantage  until  the 
pressure  reaches  150  Ibs.  The  compound  condensing  engine 
becomes  less  economical  than  the  triple-expansion  engine  for 
pressures  greater  than  150  Ibs. 

The  single-cylinder  engine,  Fig.  112,  is  more  economical  than 
the  compound  engine  when  the  number  of  expansions  of  the 
steam  is  less  than  four.  When  there  are  from  four  to  six  ex- 
pansions there  is  very  little  difference  in  the  economy.  With 
from  six  to  fifteen  expansions  the  compound  engine  is  more 
economical.  When  the  number  of  expansions  exceeds  fifteen  it 
is  usual  to  use  a  triple-expansion  engine. 


FIG.  130. — Tandem  arrangement  of  cylinders.  . 

154.  Tandem  Compound  Engines. — A  tandem  compound 
engine,  Fig.  130,  is  one  in  which  the  two  cylinerds  are  placed  one 
in  front  of  the  other.  The  pistons  of  the  two  cylinders  are 
attached  to  the  same  piston  rod,  and  there  is  but  one  connecting 
rod  and  crank.  The  steam  flows  directly  from  the  high-pressure 
cylinder  into  the  low-pressure  cylinder,  and  the  connecting 
pipes  are  relatively  small,  there  being  no  receiver  except  the 
piping  between  the  cylinders.  The  tandem  compound  engine 
occupies  less  space  than  the  cross  compound.  The  principal 
objection  to  this  form  of  engine  is  the  difficulty  of  getting  at 


COMPOUND  ENGINES 


227 


the  cylinder  which  is  nearest  the  crank-shaft.     This  is  the  earliest 
form  of  compound  engine  used. 


155.  Cross-compound  Engine. — In  the  cross-compound  engine, 
Fig.  131,  the  two  cylinders  are  placed  side  by  side,  and  each 


228  HEAT  ENGINES 

cylinder  has  its  separate  piston  rod,  connecting  rod,  and  crank. 
The  steam,  after  leaving  the  high-pressure  cylinder,  usually 
enters  a  steam  reservoir  called  a  receiver,  and  from  this  receiver 
the  low-pressure  cylinder  takes  its  steam.  The  cranks  in  a 
cross-compound  engine  are  usually  set  90°  apart,  so  that  when 
the  high-pressure  cylinder  is  at  the  beginning  of  its  stroke  the 
low-pressure  cylinder  is  at  mid-stroke.  A  cross-compound 
engine  with  cranks  at  90°  must  always  be  provided  with  a  receiver, 
as  the  low-pressure  cylinder  may  be  taking  steam  when  the 
high-pressure  cylinder  is  not  exhausting.  The  cross-compound 
engine  occupies  a  much  larger  space  than  the  tandem  engine, 
but  the  parts  are  lighter.  Each  piston,  cross-head,  connecting 
rod,  and  crank  does  only  approximately  one-half  the  work  that 
they  would  do  in  a  tandem  engine.  The  turning  effort  on  the 
crank-shaft  is  made  more  uniform  by  placing  the  crank  at  90°. 
This  reduces  the  size  of  the  fly-wheel  necessary  to  overcome 
the  fluctuation  of  the  speed  of  the  engine,  and  also  assists  the 
governing. 

A  vertical  cross-compound  engine  is  often  termed  a  "fore  and 
aft"  compound. 

156.  Ratio  of  Cylinders  in  the  Compound  Engine. — In  the 
compound  engine  the  strokes  of  the  two  cylinders  are  usually  the 
same.  If  we  represent  the  ratio  of  the  volumes  of  the  two 
cylinders  by  L,  and  the  diameter  of  the  high-pressure  cylinder 
by  d,  and  that  of  the  low-pressure  cylinder  by  D,  then 

*  -  f 

The  value  of  L  should  be  such  as  to  avoid  a  fall  in  pressure, 
termed  "drop,"  between  the  exhaust  pressure  in  the  high-pres- 
sure cylinder  and  the  admission  pressure  in  the  low-pressure 
cylinder.  The  value  of  L  varies  from  2J  to  4  for  automatic 
high-speed  engines,  and  from  3  to  4|  for  engines  of  the  Corliss 
type.  L  is  equal  to  the  quotient  of  the  number  of  times  the  steam 
is  expanded  in  the  engine  divided  by  the  number  of  expansions  in 
the  high-pressure  cylinder. 

The  ratio  of  expansion,  r,  in  a  compound*  engine  is  equal  to 
the  ratio  of  the  total  volume  of  the  low-pressure  cylinder,  or  cylinders, 
to  that  of  the  high  up  to  the  point  of  cut-off.  That  is,  it  is,  as  in 

*  See  note  at  bottom  of  page  225. 


COMPOUND  ENGINES  229 

the  case  of  a  single-cylinder  engine,  the  ratio  of  the  final  to  the 
initial  volume  occupied  by  the  steam  while  in  the  engine. 

This  ratio,  r,  may  be  varied  in  an  engine  by  varying  the  point 
of  cut-off  in  the  high-pressure  cylinder.  It  is  customary  to 
proportion  an  engine  and  so  set  the  valves  that  each  cylinder 
does  an  equal  amount  of  work.  This,  however,  is  not  always 
the  case,  some,  engines  being  designed  to  give  equal  ranges  of 
temperatures  in  the  cylinders.  Theoretically  this  gives  the  best 
economy. 

The  proportion  of  work  that  is  done  by  each  cylinder  may  be 
adjusted  by  changing  the  low-pressure  cut-off.  The  shorter  the 
cut-off  in  the  low-pressure  cylinder,  the  less  the  steam  taken 
from  the  receiver  and  the  higher  the  pressure  in  the  receiver. 
Increasing  the  pressure  in  the  receiver  causes  a  higher  back 
pressure  for  the  high-pressure  cylinder,  and  consequently  less 
work  done  by  that  cylinder.  Increasing  the  low-pressure  cut-off 
will  decrease  the  work  done  by  the  low-pressure  cylinder.  Theo- 
retically, changing  the  cut-off  in  the  low-pressure  cylinder  does  not 
change  the  gross  horse-power  developed  by  the  engine,  but  in  actual 
practice  this  does  not  hold  absolutely  true,  although  the  change 
is  very  slight.  The  equalization  of  the  work  in  the  two  cylinders 
cannot  be  accomplished  in  most  engines,  as  in  equalizing  the 
work  at  different  loads  an  excessive  drop  may  be  produced  be- 
tween the  cylinders. 

157.  Horse-power  of  a  Compound*  Engine. — In  determining 
the  horse-power  of  a  compound*  engine  from  the  indicator  cards, 
the  card  from  each  end  of  each  cylinder  is  worked  up  and  the 
horse-power  calculated  for  each,  and  the  sum  of  the  horse-powers 
determined  from  each  card  will  be  the  horse-power  of  the  engine. 

In  determining  the  horse-power  that  a  compound*  engine 
ought  to  develop  it  is  necessary  to  know  the  absolute  initial 
steam  pressure,  the  total  number  of  expansions  of  steam,  the 
number  of  strokes  per  minute,  the  length  of  the  stroke,  and  the 
diameter  of  the  high-  and  low-pressure  cylinders. 

The  horse-power  is  then  determined  as  though  there  were  but  one 
cylinder,  and  that  one  the  size  of  the  low-pressure  cylinder,  and  the 
total  expansion  of  steam  took  place  in  that  cylinder.  The  reason 
for  this  is  apparent  when  we  consider  that  the  power  of  any  engine 
per  stroke  depends  on  the  weight  of  steam  admitted  and  its 

*  See  note  at  bottom  <5T  page  225. 


230  HEAT  ENGINES 

ratio  of  expansion,  and  that  all  the  power  of  the  compound* 
engine  could  be  developed  in  its  low-pressure  cylinder  if  we  ad- 
mitted into  that  cylinder  the  same  weight  of  steam  as  was  ad- 
mitted to  the  high-pressure  cylinder,  expanded  the  steam  in  this 
cylinder  the  same  number  of  times  as  it  was  expanded  in  the 
whole  engine,  and  exhausted  against  the  same  back  pressure. 
If  the  horse-power  obtained  by  assuming  all  the  work  done  in 
the  low-pressure  cylinder  be  multiplied  by  a  card  factor,  the 
result  will  be  equal  to  the  horse-power  of  the  engine.  This 
may  be  expressed  mathematically  as  follows: 

Let   D  =  the  diameter  of  the  low-pressure  cylinder. 

d  =  the  diameter  of  the  high-pressure  cylinder. 
A  =  the  area  of  the  low-pressure  cylinder  in  square  inches. 

/  =  the  length  of  stroke  of  the  engine  in  feet. 

p  =  the  mean  effective  pressure  for  the  whole  engine. 

n  =  number  of  revolutions  per  minute. 

x  =  the  per  cent,  of  the  stroke  to  the  point  of  cut-off  in 
the  high-pressure  cylinder. 

T  =  ratio  of  expansion  for  the  whole  engine. 

e  =  the  card  factor. 

Pi  =  initial  pressure  steam  entering  the  engine. 
pz  =  pressure  of  the  exhaust. 

D2 
Then  r  -    ~2  =  (2) 


Pi  (1  +  logs)  . 

and  p  =  e  —  -  p2  (3) 


2plAn 
Horse-power  .  • 


The  value  of  the  factor  e  depends  upon  the  type  of  the  engine, 
and  varies  from  .70  to  .80  for  automatic  high-speed  engines,  and 
from  .75  to  .85  for  a  Corliss  engine. 

^        ^          ^       t> 
Example.—  A    15"  X  24"  X  36"  X  30"    engine    runs     100    r.p.m. 

Cut-off  in  the  H.P.  cylinder,  f  stroke;  in  the  intermediate  cylinder, 
f  stroke;  in  the  L.P.   cylinder,  \  stroke.     Steam  pressure,  225  Ibs. 

*  See  note  at  bottom  of  page  225. 


COMPOUND  ENGINES 


231 


Engine  exhausts  into  a  condenser  having  a  vacuum  of  26  in.  Ba- 
rometer reading,  28.65  in.  Assume  a  card  factor  of  .80. 

Indicator  cards  were  taken  from  the  engine  with  the  following  areas: 
H.P.  cylinder,  head  end,  1.32  sq.  in.,  crank  end,  1.35  sq.  in. ;  intermediate 
cylinder,  head  end,  1.8  sq.  in.,  crank  end,  1.71  sq.  in.;  L.P.  cylinder, 
head  end,  2.01  sq.  in.,  crank  end,  2.04  sq.  in.  Length  of  all  cards,  3  in. 
A  160  Ib.  spring  was  used  on  the  H.P.  cylinder,  a  50  Ib.  spring  on  the 
intermediate,  and  a  20  Ib.  spring  on  the  L.P.  The  diameters  of  the 
piston  rods  were  as  follows:  H.P.  cylinder  2  in.;  intermediate  cylinder, 
2^  in.;  L.P.  cylinder,  3  in. 

(a)  What  is  the  rated  H.P.  of  the  engine? 

(6)  What  per  cent,  of  the  rated  H.P.  is  being  developed? 

Solution. — (a) 

Atmospheric  pressure 

=  28.65  X  .491  =  14  Ibs. 

Exhaust  pressure,  p2,  =  (28.65  -  26)  X  .491  =  1.3  Ibs. 
D2  36  X  36          8  X  36  X  36 


r  = 


xd*       f  X  15  X  15       3  X  15  X  15 


=  15.35. 


M.E.P   =  e 


+  log<r)  -  pt  \  . 


239 


M.E.P.  =  .8  {    """    (1  +  log-15.35)  -  1.3  \  =  .8(58.1  -  1.3) 

[  lO.oO  J 

=  45.46  Ibs. 

Area  L.P.  cylinder  =  3.1416  X  18  X18  =  1018  sq.  in. 
2  X  45.46  X  2.5  X  1018  X  100 


Rated  I.H.P. 


Practically  a  700-H.P.  engine. 


33000 


H.P.,  H.E.,     3 


=  70.3  Ibs. 


* 


M.E.P. 


H.P.,  C.E.,     3     X  160  =  72  " 

I  M.P.,  H.E.,     Q8  X    50  =  30  " 
6 

M.P.,  C.E.,  1Q71  X    50  =  28.5  " 

o 

L.P.,  H.E.,  ~  l  X    20  =  13.4  " 


V 


L.P.,  C.E., 


o 


X    20  =  13.6    " 


232 


HEAT  ENGINES 


Area 


7T7.52 

7r(7.52  - 

I2) 

=      176.7  sq.  in. 
=     173.6  "   " 

7T122 

=     452 

n  a 

7r(122  - 

1.252) 

=     447 

a     i( 

7T182 

=   1018 

a    (i 

7r(182  - 

1.52) 

=    1011 

ii    (i 

I.H.P. 


LN        2.5  X  100 
Constant  =  =  ~  =  .007575. 


H.P.,  H.E. 
H.P.,  C.E. 
M.P.,  H.E. 
M.P.,  C.E. 
L.P.,  H.E. 
L.P.,  C.E. 


70.3  X  176.7  X  .007575  =  94.4 
72  X  173.6  X  .007575  =  94.7 
30  X  452  X  .007575  =  102.5 

28.5  X  447  X  .007575  =  96.5 

13.4  X  1018  X  .007575  =  103.5 

13.6  X  1011  X  .007575  =  104.2 

Total  =  595.8 


Per  cent,  of  rated  H.P.  developed 
595.8 


700 


=  .851  =  85.1  per  cent. 

f(a)    700  H.P. 
1(6)  85.1  per  cent. 

158.  Combined  Indicator  Cards. — The  combined  diagram  is 
a  hypothetical  figure  which  would  be  obtained  if  the  admission, 
expansion  and  exhaust  all  took  place  in  one  cylinder,  and  that 
the  low-pressure  cylinder  of  a  compound*  engine.  It  is  a  dia- 
gram on  which  may  be  measured  the  pressure  of  the  steam  at 
any  point  in  the  stroke  of  any  of  the  cylinders,  and  the  volume 
of  that  steam.  In  it  the  indicator  card  from  each  cylinder 
appears  in  its  true  proportion. 

When  combining  the  cards  from  a  compound*  engine,  it  is  first 
necessary  to. reduce  them  all  to  the  same  scale  of  pressures  and 
volumes.  It  is  generally  more  convenient  to  use  the  diagram 
from  the  low-pressure  cylinder  as  the  basis  to  which  to  change  the 
other  diagrams. 

On  each  of  the  diagrams  lay  off  a  vertical  line  back  of  the  ad- 
mission line  a  distance  equal  to  the  clearance  volume  for  that  par- 
ticular cylinder.  These  lines  represent  the  lines  of  zero  volume. 

Divide  the  high  and  intermediate  diagrams  into  any  convenient 

*  See  note  at  bottom  of  page  225. 


COMPOUND  ENGINES 


233 


number  of  parts  by  vertical    lines    spaced    equidistant    apart. 
Multiply  the  distance  from  the  atmospheric  line  of  each  of  the 


i  5 


points  where  these  vertical  lines  cross  the  diagrams,  by  the  ratio 
of  the  scale  of  the  spring  used  in  the  cylinder  being  considered  to 


234  HEAT  ENGINES 

that  used  in  the  low-pressure  cylinder.  The  results  will  be  the 
ordinates  of  the  points  to  be  plotted  when  drawing  the  combined 
diagram.  (The  pressure  in  each  cylinder  at  any  point  in  the 
stroke  may  be  determined  from  the  indicator  card  for  that 
cylinder,  knowing  the  value  of  the  spring  used  and  the  position 
of  the  atmospheric  line).  Multiply  the  horizontal  distances  from 
the  zero  volume  line  to  the  points  of  intersection  of  the  vertical 
lines  and  the  diagram,  by  the  ratio  of  the  volume  of  the  cylinder 
under  consideration  to  the  volume  of  the  low-pressure  cylinder. 
The  results  obtained  will  be  the  abscissae  of  the  points  on  the 
combined  diagram. 

Now  plot  the  results  using  the  atmospheric  line  and  line  of  zero 
volume  on  the  low-pressure  diagram  as  the  horizontal  and  vertical 
axes  from  which  to  measure  the  ordinates  and  abscissae.  Through 
the  points  so  plotted,  draw  the  combined  diagram. 

Fig.  132  shows  the  combined  diagram  from  a  triple-expansion 
pumping  engine.  The  indicator  cards  for  the  high,  intermediate, 
and  low-pressure  cylinders,  have  all  been  reduced  to  the  same 
scale  of  volumes  and  pressures.  The  ordinates  in  this  diagram  are 
absolute  pressures  and  the  abscissae  are  volumes.  The  indicator 
card  from  each  cylinder  was  divided  into  an  equal  number  of 
parts  and  the  pressure  and  volume  at  each  of  these  points  was 
computed  and  plotted  in  the  figure.  The  indicator  card  for  each 
cylinder,  it  will  be  noticed,  does  not  begin  at  the  zero  volume 
line,  the  difference  between  zero  volume  of  the  indicator  card 
and  the  zero  of  volumes  representing  the  volume  of  the  clear- 
ance. The  dotted  saturation  curve  shows  what  the  curve  of 
expansion  would  have  been  if  the  actual  weight  of  steam  ex- 
panding in  the  engine  had  remained  saturated. 

PROBLEMS 

1.  A  compound  engine  is  8"  X  16"  X  12"  and  runs  300  r.p.m.  Initial 
steam  pressure,  150  Ibs.  absolute;  back  pressure,  2  Ibs.  absolute.  Cut-off 
in  high-pressure  cylinder  at  \  stroke.  If  the  steam  expands  along  an 
isothermal  of  a  perfect  gas  and  the  card  factor  is  70  per  cent.,  what  I.H.P. 
will  the  engine  develop? 

v  2.  A  single-acting  compound  engine  is  9"  X  15"  X  9";  initial  pressure, 

125  Ib.  gage;  back  pressure,  atmospheric;  cut-off  in  high-pressure  cylinder, 
i  stroke;  r.p.m.,  250.  Assume  card  factor  of  80  per  cent.  What  would  be 
the  horse-power  rating  of  the  engine? 

3.  A  compound  engine  is  27"  X  35"  X  48"  and  runs  80  r.p.m.  Initial 
pressure,  125  Ibs.;  back  pressure,  2  Ibs.  absolute;  cut-off  in  the  high-pressure 


COMPOUND  ENGINES  235 

cylinder,  £  stroke;  card  factor,  80  per  cent.  What  will  be  its  horse-power 
if  each  cylinder  develops  an  equal  number  of  horse-power? 

4.  A  triple-expansion  engine  is  20"  X  27"  X  40"  X  36".  Cut-off  in  high- 
pressure  cylinder,  \  stroke;  in  intermediate  cylinder,  \  stroke;  and  in  low 
pressure  cylinder,  \  stroke.  Steam  pressure,  135  Ibs.;  back  pressure,  2  Ibs. 
absolute.  Engine  is  double  acting  and  runs  50  r.p.m.  Assuming  a  card 
factor  of  80  per  cent.,  what  is  the  rated  horse-power  of  the  engine? 

6.  A  city  pumping  engine  is  27"  X  35"  X  55"  X  48";  cut-off  in  high- 
pressure  cylinder,  \  stroke.  Steam  pressure,  130  Ibs.;  back  pressure,  2  Ibs. 
absoluta;  r.p.m.,  40.  Assume  card  factor  of  85  per  cent.  What  is  its  rated 
horse-power? 

6.  An  engine  is  9"  X  15"  X  9"  and  runs  320  r.p.m.,  the  cut-off  in  the 
high-pressure  cylinder  being  \  stroke.     Steam  pressure,  125  Ibs.;  back  pres- 
sure, 3  Ibs.  absolute.     Engine  is  single  acting  and  the  area  of  the  indicator 
card  from  high-pressure  cylinder  is  .9  sq.  in.;  from  the  low-pressure  cylinder, 
9.  sq.  in.     The  length  of  each  is  2.35  in.     An  83-lb.  spring  is  used  in  indicator 
on  high-pressure  cylinder  and  a  40-lb.  spring  in  indicator  on  low-pressure 
cylinder.     The  engine  is  fitted  with  a  Pony  brake  carrying  a  gross  weight 
of  120  Ibs.     The  tare  of  the  brake  is  20  Ibs.  and  the  length  of  the  brake  arm 
is  51  in.     Find  the  I.H.P.;  B.H.P.;  F.H.P.;  and  mechanical  efficiency. 

7.  A  triple-expansion  engine  is  18"  X  28"  X  36"  X  24"  and  runs  180 
r.p.m.  Steam  pressure,  200  Ibs.  absolute;  vacuum,  28  in.     Diameter  of 
piston  rod  for  H.P.  cylinder,  2  in.;  for  M.P.  cylinder,  3  in.;  for  L.P.  cylinder, 
4  in.     Indicator  spring  for  H.P.  cylinder,  160  Ibs.;  for  M.P.  cylinder,  60 
Ibs;  for  L.P.  cylinder,  20  Ibs.     The  area  and  lengths  of  indicator  cards  are 
as  follows: 

H.E.,  H.P.  Cyl.,  area  2  square  inch,  length  3    in. 
C.E.       "         "       "     1\       "        "          "     3     " 
H.E.,  M.P.     "       "     U       "        "          "     2^   " 
C.E.       "         "       "     If       "        "          "     2J   " 
H.E.,  L.P.       "       "    3  "          "     3     " 

C.E.       "         "       "    3£       "  "     3i   " 

Find  the  total  I. H.P. 


CHAPTER  XIV 
CONDENSERS  AND  AIR  PUMPS 

159.  There  are  two  general  forms  of  condensers  in  use,  the 
jet  condenser  and  the  surface  condenser.     In  condensers  of  the 


FIG.  133.— Jet  condenser. 

jet  type,  the  condensing  water  and  the  steam  are  brought  into 
contact  with  each  other,  while  in  the  surface  condensers  the  con- 

236 


CONDENSERS  AND  AIR  PUMPS 


237 


densing  water  and  the  steam  condensed  do  not  come  in  direct 
contact. 

160.  Jet  Condensers. — There  are  two  principal  types  of  jet 
condensers,  the  regular  jet  type  and  the  barometric  condenser. 

Fig.  133  shows  a  jet  condenser  of  the  regular  type.  The 
exhaust  steam  from  the  engine  enters  at  A,  and  the  injection  water 
at  B.  These  are  mixed  in  the  combining  chamber  F.  The  water 
in  entering  the  combining  chamber  is  sprayed  by  a  rose-head  on 
the  end  of  the  inj ection  pipe  at  D.  The  condensation  of  the  steam 


FIG.  134. — Barometric  condenser. 

reduces  its  volume  many  times,  and  this  reduction  of  volume 
forms  a  vacuum  in  the  chamber  F.  This  vacuum  is  maintained 
by  the  pump  G,  which  removes  the  condensing  water,  and  air 
which  is  present  in  small  quantities. 

In  the  position  shown  in  the  figure,  the  piston  is  moving  toward 
the  left,  forcing  the  water  in  the  crank  end  of  the  cylinder  out 
through  valve  I  into  the  discharge  pipe  J  and  holding  valve  H 
closed.  At  the  same  time  the  suction  in  the  head  end  of  the 
cylinder  closes  valve  /'  and  opens  valve  Hf  drawing  in  the  water 
and  air  from  the  combining  chamber  F.  If  the  source  of  con- 
densing water  is  not  more  than  15  ft.  below  point  B,  it  is  pos- 


238 


HEAT  ENGINES 


sible  to  draw  the  water  into  the  condenser  by  the  vacuum  in  the 
chamber  F. 

In  the  barometric  condenser,  Figs.  134  and  135,  the  water  and 
steam,  after  coming  in  contact,  pass  as  water  through  a  narrow 
.opening  in  the  throat  of  the  condenser.  The  water  passing 
through  this  narrow  throat  carries  the  air  with  it.  The  condensa- 
tion of  the  steam  forms  a  vacuum,  but  the  condensing  water  must 
be  pumped  into  the  combining  chamber,  as  this  chamber  is  ele- 


Relief  Valve 


FIG.   135. — Complete  installation  of  barometric  condenser. 

vated  at  least  35  ft.  above  the  hot  well,  so  that  the  pressure  of  the 
atmosphere  upon  the  surface  of  the  water  in  this  well  cannot  force 
it  up  into  the  chamber.  The  quantity  of  water  passing  through 
the  throat  of  the  condenser  cannot  be  decreased  very  much,  as, 
if  the  velocity  of  the  water  passing  is  reduced  to  any  extent,  it  will 
be  insufficient  to  maintain  the  vacuum.  For  this  reason  the 
barometric  condenser  is  not  adapted  to  variable  loads. 

The  jet  condenser  is  the  form  most  used  in  stationary  plants, 


CONDENSERS  AND  AIR  PUMPS 


239 


240  HEAT  ENGINES 

as  it  is  less  expensive  to  install,  requires  less  repairs,  and,  where 
clean  water  is  available,  gives  as  good  results  as  the  surface 
condenser. 

161.  Surface  Condensers. — In  a  surface  condenser,  Fig.  136, 
'the  steam  to  be  condensed  and  the  cooling  water  do  not  come  in 

direct  contact  with  each  other.  The  cooling  water  is  circulated 
on  one  side  of  a  series  of  tubes,  and  the  steam  is  condensed  by 
coming  in  contact  with  the  other  side  of  the  tubes.  The  con- 
densed steam  is  drawn  off  by  the  air  pump.  The  condensing 
water  is  drawn  or  forced  through  by  the  circulating  pump. 
The  surface  of  the  tubes  which  come  in  contact  with  the  steam 
is  the  condensing  surface.  The  tubes  are  always  of  small  diam- 
eter, and  the  metal  is  made  as  thin  as  possible  and  usually  of 
brass.  A  surface  condenser  is  used  where  the  cooling  water  is  not 
suitable  for  feed  water,  and  it  is  necessary  to  use  the  same  water 
over  and  over  again  for  making  steam  in  the  boilers.  The  con- 
densed steam  being  distilled  water,  contains  no  scale-forming 
matter  and  is  excellent  for  feed  water.  Care  must  be  taken, 
however,  to  see  that  none  of  the  oil  contained  in  the  exhaust  steam 
is  allowed  to  go  back  to  the  boiler.  With  the  surface  condenser, 
the  nature  of  the  cooling  water  is  immaterial  as  none  of  it  will  be 
used  as  feed  water.  This  is  the  form  of  condenser  always  used  in 
salt  water,  marine  practice. 

162.  Air  Pumps. — The  air  pump  is  sometimes  operated  directly 
from  the  engine.     This  is  done  to  avoid  the  use  of  steam  by  the 
independent  condenser  pump,   which  is  always  uneconomical, 
using  from  70  to  120  Ibs.  of  steam  per  I.H.P.  per  hour. 

The  water  and  air  are  usually  discharged  through  the  discharge 
pipe  into  a  tank,  or  well,  called  the  hot  well,  and  the  overflow 
from  this  hot  well  runs  into  the  sewer,  or  river.  The  feed  water 
for  the  boilers  in  a  condensing  plant  is  usually  taken  from  the  hot 
well. 

In  many  plants,  when  very  high  vacuum  is  desired,  there  is 
added  a  dry  air  pump,  in  addition  to  the  devices  already  described. 
This  is  attached  to  the  combining  chamber  so  as  to  remove  the  air 
from  it,  the  regular  vacuum  pump  removing  only  the  water. 
Vacuums  as  high  as  28  in.  and  over  are  maintained  in  these  plants. 

163.  Amount  of  Cooling  Water. — The  amount  of  cooling  water 
required  in  a  condenser  depends  upon  the  temperature  of  the 
water  and  the  degree  of  vacuum  desired.     If  the  temperature  in 
the  condenser  is  too  high,  low  vacuum  cannot  be  obtained,  as  the 


CONDENSERS  AND  AIR  PUMPS  241 

pressure  in  the  condenser  cannot  be  less  than  the  pressure  cor- 
responding to  the  temperature  of  the  boiling  point  in  the  conden- 
ser. If  the  temperature  of  the  water  in  the  hot  well  is  120°, 
the  corresponding  pressure  as  given  in  the  steam  tables  is  1  .  7  Ibs., 
which  is  the  lowest  possible  pressure  that  can  be  obtained.  If 
a  lower  vacuum  is  desired,  the  temperature  of  the  water  leaving 
the  condenser  must  be  lowered.  This  can  be  done  in  two  ways, 
by  increasing  the  amount,  or  decreasing  the  temperature,  of  the 
cooling  water. 

If  we  let   ti  =  the  initial  temperature  of  the  cooling  water; 

t2  =  the  final  temperature  of  the  cooling  water  (and, 
in  a  jet  condenser,  of    the   condensed  steam 
also)  ; 
£3  =  the  temperature  of  the  condensed  steam  leav- 

ing a  surface  condenser; 
H  =  heat  (above  32°)  in  the  steam  entering  the  con- 

denser; 
W  =  the  weight  of  the  cooling  water  entering  per 

minute  ; 
w  =  the  weight  of  steam   condensed  per  minute; 

then  the  heat  given  up  by  the  steam  in  a  jet  condenser 

=  win  -  (t2  -  32)!,  (i) 

and  the  heat  received  by  the  water 

=  W(t2  -  t,).  (2) 

But  these  two  expressions  must  be  equal,  and  equating  and  solv- 
ing, 


W  -  (3) 

The  amount  of  cooling  water  per  pound  of  steam  entering  a  sur- 
face condenser  is  larger  than  that  used  in  a  jet  condenser,  as  the 
temperature  of  the  condensed  steam  is  higher  than  the  tempera- 
ture of  the  cooling  water  leaving  the  condenser.  Substituting 
/3  for  /2  in  equation  (1),  equation  (3)  becomes, 

w  .  yiH^JS)± 
[2  —  h 

which  is  the  expression  for  the  weight  of  cooling  water  used  to 
condense  w  pounds  of  steam  per  minute  in  a  surface  condenser. 

16 


242  HEAT  ENGINES 

In  ordinary  stationary  practice,  1  sq.  ft.  of  cooling  sur- 
face is  allowed  for  every  10  Ibs.  of  steam  condensed  per  hour, 
except  in  the  case  of  turbines  using  high  vacuum,  where  1  sq.  ft. 
is  allowed  for  every  4  to  8  Ibs.  of  steam.  In  navy  practice,  from 
1  to  1  J  sq.  ft.  of  surface  are  allowed  for  every  indicated  horse-power. 

164.  Increase  of  Power  by  Use  of  Condenser.  —  Condensing  the 
exhaust  steam  diminishes  the  back  pressure  by  creating  a  partial 
vacuum  in  the  exhaust  system.  This  vacuum  is  generally 
measured  in  inches  of  mercury.  It  is  seldom  that  the  vacuum 
maintained  in  the  condenser  exceeds  26  in.,  and  24  in.  is  more 
common.  In  the  expression  for  mean  effective  pressure, 


M.E.P.  -  .      Pl 

the  quantity  affected  by  the  vacuum  is  the  term  p2.  In  a  non- 
condensing  engine  this  is  usually  about  15  Ibs.  and  the  M.E.P. 
about  40  Ibs.,  but  in  the  condensing  engine  the  effect  of  adding 
a  condenser  is  to  lower  pz  to  about  2  Ibs.,  and  increase  the  M.E.P. 
for  a  single  cylinder  engine  to  53  Ibs.,  adding  to  the  horse-power 
of  the  engine  about  20  per  cent. 

165.  Condensers  for  Steam  Turbines.  —  In  most  steam  turbine 
plants,  surface  condensers  are  used,  principally  for  the  reason  that 
the  exhaust  from  the  steam  turbine  does  not  contain  oil,  and  when 
condensed  is  an  ideal  feed  water,  as  it  contains  no  scale-producing 
matter.  It  is  also  possible  in  a  surface  condenser  to  use  very 
large  quantities  of  circulating  water,  and  thus  reduce  the  tempera- 
ture of  the  condenser. 

In  turbine  plants  an  increase  in  the  vacuum  increases  the  econ- 
omy of  the  turbine  materially,  and  every  means  is  used  to  get  the 
highest  possible  vacuum. 

PROBLEMS 

1.  A  150  H.P.  engine  has  a  guaranteed  steam  consumption  of  20  Ibs.  per 
I.H.P.  per  hour;     On  being  tested,  it  was  found  that  it  took  21.6  Ibs.     The 
engine  operates    10  hours  per  day,  300   days   per   year,    and    cost   $4500 
to   install.     The  steam  costs   25   cents  per    1000  Ibs.   to  produce.  '   How 
much  should  be  deducted  from  the  cost  price  to  compensate  the  purchaser 
for  the  increased  cost  of  operation  above  that  required  under  the  guarantee? 
Allow  6  per  cent,  interest  and  5  per  cent,  depreciation.     (Suggestion:   Find 
the  present  worth  of  an  annuity  equal  to  the  loss  per  year  due  to  the  excess 
steam  consumption.) 

2.  Given  a  plant  equipped  with  two  5000  H.P.  engines  that  use  20  Ibs. 
of  steam  per  horse-power  per  hour.     Feed  temperature,  70°;  steam  pressure, 


CONDENSERS  AND  AIR  PUMPS  243 

150  Ibs.  Boilers  evaporate  9  Ibs.  of  water  per  pound  of  coal.  Coal  cost 
$2.25  per  ton.  Engines  run  10  hours  a  day,  300  days  in  the  year.  If 
these  engines  are  taken  out  and  sold  for  $5  per  horse-power  and  new  ones 
using  only  12  Ibs.  of  steam  per  horse-power  per  hour  are  installed;  (a)  how 
many  boiler  horse-power  will  be  saved;  (6)  how  much  will  be  saved  per  year 
on  the  coal  bill;  (c)  how  much  can  be  paid  for  the  new  engines  if  they  are  to 
return  6  per  cent,  on  the  investment,  the  depreciation  of  the  engines  being 
4  per  cent.  ? 

3.  A  100  H.P.  automatic  engine  uses  32  Ibs.  of  steam  per  I.H.P.  per  hour 
and  costs  $1500.     A  100  H.P.  Corliss  engine  uses  26  Ibs.  of  steam  per  I.H.P. 
per  hour  and  costs  $2200.     Steam  in  the  plant  costs  20  cents  per  1000  Ibs. 
The  plant  runs  10  hours  per  day  and  300  days  per  year.     Allowing  5  per 
cent,   interest,    10  per     cent,  depreciation  on  the  high-speed  engine,  and 
7  per  cent,  on  the  Corliss,  (a)  which  will  be  the  most  economical  engine  to 
biiy?     (6)  How  much  will  the  one  engine  save  over  the  other  per  year  in 
operation? 

4.  A  power  plant  is  to  deliver  1000  K.W.  at  the  switch  board.     A  steam 
engine  can  be  installed  which  will  give  an  economy  of  14^  Ibs.  steam  per 
I.H.P.  per  hour.     Steam  pressure,  145  Ibs.;  feed  water,  150°;  dry  steam. 
Engine  efficiency,  92  per  cent.;  generator  efficiency,  95  per  cent.     A  steam 
turbine  can  be  installed  which  will  give  a  steam  consumption  of  20  Ibs.  per 
K.W.  per  hour.     Steam  pressure,  145  Ibs.;  150°  superheat.     Feed  water, 
150°.     Cost  of  generating  steam  in  each  case,  20  cents  per  1,000,000  heat 
units.     Which  is  the  more  economical  installation  and  how  much  is  saved 
by  this  one  per  hour  over  the  other? 


CHAPTER  XV 
STEAM  TURBINES 

166.  Historical. — From  the  earliest  time  attempts  have  been 
made  to  produce  a  rotary  motion  by  steam  without  converting 
a  reciprocating  motion  into  the  rotary  motion  Devices  for 
doing  this  have,  with  the  exception  of  the  steam  turbine,  been 
a  failure. 

The  modern  steam  turbine  is  the  revival  of  the  earliest  form 
of  steam  motor.  The  first  contrivance  of  this  kind  dates  back 
to  Hero's  turbine,  shown  in  Fig.  137,  which  was  designed  two 
centuries  before  the  birth  of  Christ.  Hero's  turbine  consisted 


M 


FIG.  137. — Hero's  turbine. 

of  a  hollow  spherical  vessel  pivoted  on  a  central  axis.  It  was 
supplied  with  steam  from  a  boiler  through  the  support  M  and 
one  of  the  pivots.  The  steam  escaped  from  the  spherical  vessel 
through  bent  pipes  or  nozzles,  N,  N,  facing  tangentially  in  oppo- 
site directions.  Rotation  was  produced  by  the  reaction  due  to  the 
steam  discharged  from  the  nozzles,  just  as  a  Barker's  mill  is 
moved  by  the  water  escaping  from  its  arms.  Hero's  turbine 
was  moved  by  the  reaction  of  the  steam  jets  alone,  so  that  it  is 
called  a  reaction  turbine. 

244 


STEAM  TURBINES 


245 


Giovanni  Branca  in  1629  designed  a  steam  turbine  as  shown 
in  Fig.  138.  Steam  issues  from  the  nozzle  in  the  mouth  of  the 
figure  in  the  form  of  a  jet.  This  jet  strikes  the  blades  of  a  wheel 
and  causes  it  to  rotate.  The  wheel  is  moved  by  the  impulse  of 
the  steam  jet  exerted  upon  the  blades  of  the  turbine  wheels. 
The  Branca  turbine  is  then  of  the  impulse  type. 

167.  Forces  of  Impulse  and  Reaction. — Fig.  139  illustrates 
the  force  of  both  impulse  and  reaction.  A  tank  filled  with 
water  is  suspended  from  above,  and  from  one  side  a  jet  of 
water  is  allowed  to  escape  through  a  nozzle.  This  issuing  jet 
impinges  against  a  block  of  wood  having  a  curved  surface 
which  turns  the  jet  of  water  back  against  its  original  direction. 


FIG.  138. — Branca's  turbine. 


The  water  striking  the  block  produces  a  force  which  tends  to 
move  the  block  in  the  original  direction  of  the  jet,  and  which 
is  the  result  of  the  jet  being  turned  in  direction  by  the  curved 
surface.  This  force  may  be  termed  an  impulsive  force.  At  the 
same  time  the  stream  issuing  from  the  tank  exerts  a  reaction 
upon  it,  tending  to  move  it  to  the  left.  This  may  be  termed  the 
force  of  reaction,  and  is  caused  by  the  fact  that  the  particles  of 
water  are  accelerated  in  velocity  from  practically  zero  inside  the 
tank  to  a  maximum  at  the  mouth  of  the  nozzle.  In  doing  this 
accelerating  a  continuous  force  must  be  impressed  upon  the  par- 
ticles, and  the  tank  and  its  contents  must  at  the  same  time, 
sustain  a  reactive  force  of  the  same  amount.  This  reactive  force 
corresponds  to  the  recoil  or  kick  of  a  gun.  In  the  case  of  the  gun, 
however,  the  discharge  is  irregular;  while  in  the  case  of  the  nozzle 
it  is  continuous.  Reactive  force  accompanies  the  generation 
of  velocity;  and  impulsive  force  accompanies  destruction  of  it, 
or  the  deflection  of  a  jet.  Either  one  or  both  of  the  forces  above 


246 


HEAT  ENGINES 


described  may  be  used  to  give  the  driving  impetus  in  a  steam 
turbine. 

168.  Classification. — The  two  types  of  turbines  described  in 
paragraph  166  are  typical  of  the  modern  classification  of  tur- 
bines, viz: 

Impulse  turbines,  in  which  the  expansion  of  the  steam  takes 
place  only  in  the  stationary  nozzles  or  guide  vane  passages,  the 
pressure  on  both  sides  of  the  moving  blades  or  wheel  being  the 
same,  and 


FIG.  139. — Diagram  illustrating  forces  of  impulse  and  reaction. 

Reaction  turbines,  in  which  the  expansion  takes  place  either 
partially  or  completely,  in  the  moving  blades  or  buckets. 

In  Fig.  140  is  shown  an  actual  impulse  turbine  nozzle  and 
blades.  In  this  wheel,  the  motion  is  produced  by  the  impulsive 
action  of  the  steam  striking  the  blades.  The  entire  expansion  of 
the  steam  takes  place  in  the  nozzle  there  being  no  expansion  in 
the  blades. 

The  blades  and  nozzle  of  a  partial  reaction  turbine  are  shown  in 
Fig.  141.  Here  the  expansion  of  the  steam  is  not  complete  in  the 
nozzle,  which  is  so  designed  that  it  cannot  fully  expand  the  steam 
by  the  time  the  jet  has  left  it.  The  remainder  of  the  expansion  is 
completed  in  the  blade  passages.  The  nozzles  are  shaped  like 
the  blades,  and  are  called  stationary  blades.  But  they  expand 
the  steam  as  nozzles  would,  and  are  therefore  essentially  equiva- 
lent to  nozzles.  The  passageways  through  the  blades  are  conver- 


STEAM  TURBINES 


247 


gent  in  cross-sections.    In  commercial  impulse  turbines  blade  pas- 
sages are  usually  made  slightly  divergent. 

169.  Action  of  Steam  in  Turbine. — The  steam  turbine  uses 
steam  in  a  manner  fundamentally  different  from  that  in  which 
the  steam  reciprocating  engine  uses  it.  The  purpose  of  both 
machines  is  to  convert  the  potential  or  pressure  energy  of  the 
steam  into  mechanical  work.  The  reciprocating  engine  ac- 
complishes this  by  allowing  the  steam  to  exert  a  pressure 


FIG.  140. — Impulse  type.  FIG.  141. — Reaction  type. 

Turbine  nozzle  and  blades. 


directly  upon  its  piston.  In  the  steam  turbine,  there  are  two 
steps  in  the  transformation.  First,  the  potential  energy  of 
the  steam  is  converted  into  kinetic  energy;  and  second,  the 
kinetic  energy  of  the  jet  is  changed  into  mechanical  or  useful 
work.  The  first  of  the  two  operations  is  performed  by  the  nozzles 
whose  function  it  is  to  expand  the  steam  from  one  pressure  to 
another  in  such  a  way  as  to  produce  the  maximum  velocity  of  jet 
possible,  and  at  the  same  time  to  direct  this  jet  properly  upon 
the  blades.  The  second  operation  is  accomplished  by  the  blades 
or  moving  elements  whose  functions  it  is  to  abstract  the  energy 
of  velocity  from  the  steam,  and  convert  it  into  a  useful  form. 


248 


HEAT  ENGINES 


Sometimes,  the  steam  is  allowed  to  expand  partly  in  the  blades, 
in  addition  to  the  expansion  which  has  already  taken  place  in 
the  nozzle.  When  this  is  the  case,  the  turbine  belongs  to  the 
partial  reaction,  or  Parsons  type. 

170.  Turbine  Nozzles. — As  steam  flows  through  a  nozzle 
its  pressure  is  gradually  reduced.  At  the  same  time  the 
velocity  of  the  particles  of  steam  is  correspondingly  in- 
creased. The  volume  of  the  steam  also  grows  as  the  steam 
proceeds  at  a  continually  diminishing  pressure.  The  area  of 
the  nozzle  at  any  particular  cross-section  is  dependent  upon 
the  velocity  and  volume  of  the  steam  as  well  as  upon  the 
total  weight  passing  per  second.  Since  the  weight  is  the  same  for 
all  cross  sections,  if  the  velocity  and  specific  volume  increased  at 
the  same  rate,  then  the  area  of  a  nozzle  at  all  cross-sections  would 
be  the  same  and  the  nozzle  would  be  nothing  more  than  a  tube  of 


Entrance 


Throat 


Mouth 


Throat 


FIG.    142. — Turbine   nozzle- 
longitudinal  cross-section. 


Entrance 


FIG.   143. — Ordinary   form    of 
nozzle. 


Mouth 


uniform  diameter.  As  a  matter  of  fact,  however,  the  velocity  and 
specific  volume  do  not  increase  at  the  same  rate.  During  the 
first  part  of  the  expansion  the  velocity  increases  more  rapidly 
than  the  specific  volume;  while  during  the  latter  part  of  the 
expansion  the  specific  volume  increases  more  rapidly  than  the 
velocity.  The  cross-sectional  area  of  the  nozzle  is  determined  by 
the  ratio  of  velocity  to  volume,  and  should  diminish  at  first  to  a 
minimum  value  and  then  increase  to  a  maximum  at  the  end.  The 
point  of  least  cross-section  is  called  the  throat  and  is  that  place  in 
the  steam's  progress  through  the  nozzle  at  which  the  rate  of 
increase  in  specific  volume  overtakes  the  rate  of  increase  in  ve- 
locity. Fig.  142  shows  the  general  shape  of  a  nozzle  on  a  longitu- 
dinal cross-section.  The  nozzle,  as  ordinarily  constructed  is 
shown  in  Fig.  143,  and  differs  from  the  nozzle  of  Fig.  142  in  that  the 
throat  cross-section  has  been  moved  very  near  the  entrance  end. 
The  convergent  portion  of  the  nozzle  consists  merely  of  a  rounding 
or  fillet.  The  divergent  portion  is  comparatively  long  and  of  a 
straight  taper.  All  nozzles  have  a  convergent  portion.  When 


STEAM  TURBINES 


249 


the  back  pressure  against  which  the  nozzle  is  discharging  is  58  per 
cent,  or  more  of  the  initial  pressure,  the  nozzle  should  be  wholly 
convergent  or  it  should  be  uniform  in  cross-section  from  the  throat 
to  the  discharge  end.  But  when  the  back  pressure  is  less  than 
58  per  cent,  of  the  initial  pressure  then  the  nozzle  should  have  a 
divergent  part  also.  The  size  of  the  nozzle  at  the  mouth  or  the 
relation  of  the  area  at  the  mouth  to  the  area  at  the  throat  is  a 
function  of  the  relation  between  the  back  pressure  and  the  intial 
pressure. 

The  following  table  shows  how  the  velocity  and  specific  volume 
increase  as  the  pressure  falls  during  adiabatic  expansion: 

TABLE  XXI 

RELATIVE  CHANGES  IN  VELOCITY,  SPECIFIC  VOLUME  AND  PRESSURE  OF 
STEAM  FLOWING  THROUGH  A  NOZZLE 


1 

Pressure,  Ibs.  per 
sq.  in.  abs. 

fi 

Available 
energy, 
B.T.U. 

3 

Velocity,  ft. 
per  sec. 
(theoretical) 

4 

Quality,  per 
cent. 

5 

Specific 
volume  of  dry 
steam 

6 
Actual  spec, 
vol.  (=  col. 
4  X  col.  5) 

150 

0.0 

0 

100.0 

3.012 

3.012 

135 

9.5 

689 

99.1 

3.331 

3.301 

120 

19.5                986 

98.3 

3.726 

3.663 

105 

30.0 

1225 

97.4 

4.23 

4.12 

90 

42.0 

1450 

96.3 

4.89 

4.70 

75 

56.0 

1675 

95.2 

5.81 

5.53 

60 
45 

72.5 
93.5 

1905 
2160 

93.9 
92.3 

7.17 
9.39 

6.73 
8.66 

30 

121.5 

2460 

90.2 

13.74 

12.40 

15 

168.5 

2900 

87.0 

26.27 

22.87 

171.  Speed  of  Turbine. — The  speed  at  which  the  blades  of  the 
turbine  wheel  will  give  the  best  efficiency  shows  the  velocity  at 
which  the  steam  impinges  upon  the  blade.  In  Fig.  144  the  direc- 
tion of  the  steam  jet  as  it  leaves  the  nozzle  is  represented  by  ah. 
Suppose  the  velocity  of  this  jet  to  be  2000  ft.  per  second  and  let 
the  blade  be  of  such  a  form  that  the  jet  is  turned  any  direction 
through  an  angle  of  180°.  Assuming  the  blade  to  be  stationary, 
that  is  the  turbine  wheel  blocked,  then  the  speed  of  the  steam 
relative  to  the  blade  will  be  2000  ft.  per  second  as  it  enters,  or  it 
is  the  same  as  the  absolute  velocity  with  which  the  steam  left 
the  nozzle.  Assuming  the  blade  to  be  frictionless  the  steam  will 
be  gradually  deflected  in  direction  without  loss  of  velocity  and  will 
emerge  with  the  same  speed  at  which  it  entered,  namely,  2000 
ft.  per  second.  An  impulsive  force  will  have  been  exerted  upon 


250  HEAT  ENGINES 

the  blade,  but  no  energy  will  have  been  taken  from  the  steam  jet 
because  there  has  been  no  motion  of  the  blade.  Now  assume 
another  case  in  which  the  blade  moves  with  a  speed  of  500  ft.  per 
second.  Then  the  speed  of  the  jet  as  it  enters  the  blade  will  be 
only  1500  ft.  per  second  relative  to  the  blade,  although  the  abso- 
lute velocity,  that  is  the  velocity  at  which  the  steam  leaves  the 
nozzle,  is  still  2000  ft.  per  second.  The  jet  will  continue  along 
the  blade  as  before  and  will  emerge  from  it  with  the  same  speed 


2000 


Steam  Jet 


2000 ^ 

_GOO 
1000 


2000 


0 

FIG.  144. — Diagram  showing  relative  velocities  of  steam  jet  and 


g  rea 


with  which  it  entered,  that  is,  2000  ft.  per  second.  But  since 
the  blade  is  traveling  at  500  ft.  per  second  in  a  direction  opposite 
to  that  of  the  leading  jet,  the  actual  absolute  velocity  of  the  jet 
at  exhaust  from  the  blade  will  be  only  1000  ft.  per  second.  All 
the  energy  is  not  abstracted  from  the  jet  because  the  steam  still 
has  left  an  amount  of  energy  represented  by  the  velocity  of  1000 
ft.  per  second.  It  is  evident  then  that  in  order  to  abstract  all 
the  energy  from  the  jet  it  will  be  necessary  for  the  steam  to  leave 
the  blade  with  no  velocity. 

Let  us  assume  another  case  in  which  the  blade  speed  is  1000  ft. 
per  second.  Then,  reasoning  as  before,  the  velocity  of  the  jet 
as  [it  enters  the  blade  will  be  1000  ft.  per  second  relative  to  the 
blade  and  in  this  case  the  steam  will  leave  the  blade  with  a 
relative  velocity  of  1000  ft.  per  second.  But  since  the  blade 
itself  is  now  traveling  1000  ft.  per  second,  the  steam  will  be  dis- 


STEAM  TURBINES  251 

charged  having  no  absolute  velocity.  In  this  case,  then,  all  the 
kinetic  energy  of  the  steam  has  been  abstracted  by  the  blade  and 
it  is,  therefore,  operating  under  the  most  efficient  speed.  It 
may  therefore  be  said  in  general  that  for  turbines  with  a  single 
row  of  blades  and  excluding  friction,  the  peripheral  speed  of  the 
blades  for  greatest  efficiency  should  be  one-half  the  speed  of  the 
jet.  In  actual  practice  the  steam  cannot  usually  be  deflected 
through  an  angle  of  180°  nor  can  the  blade  be  so  designed  as  to 
avoid  friction.  Both  these  causes  tend  toward  reducing  the 
speed  of  the  blade  below  one-half  the  speed  of  the  jet  for  maximum 
efficiency,  this  speed  reduction  being  from  10  to  15  per  cent. 

In  a  nozzle  expanding  steam  under  a  pressure  of  160  Ibs. 
absolute  to  a  pressure  of  1  Ib.  absolute,  the  theoretical  velocity 
of  the  issuing  jet  may  be  as  high  as  4000  ft.  per  second.  This 
would  give  a  blade  speed  of  from  1700  to  1000  ft.  per  second, 
which  is  almost  prohibitive  even  under  the  best  mechanical 
construction.  It  is  this  condition  that  has  led  to  thex develop- 
ment of  the  multiple  pressure  stage  and  multiple  velocity  stage 
turbines,  which  are  described  later  and  whose  primary  purpose 
is  the  attainment  of  a  high  efficiency  with  a  much  lower  blade 
speed. 

172.  De  Laval  Turbines. — The  De  Laval  single-stage  turbine 
consists  of  a  group  of  nozzles  located  around  the  periphery  of  a 
wheel  to  which  the  blades  are  attached.  The  blades  and  jets 
of  this  form  of  turbine  are  shown  diagrammatically  in  Fig.  145, 
and  a  plan  of  the  turbine  together  with  its  gearing  is  shown  in 
Fig.  146.  The  turbine  wheel  W  is  supported  upon  a  light  flexible 
shaft  between  the  bearing  Z,  provided  with  a  spherical  seat, 
and  a  gland  or  stuffing-box  P.  The  purpose  of  this  flexible  shaft 
is  to  permit  the  wheel,  when  running  at  high  speeds,  to  revolve 
about  its  center  of  gravity  instead  of  its  geometrical  center, 
thus  reducing  vibration  and  wear.  Teeth  are  cut  into  the  metal 
of  the  shaft  to  make  the  pinions  on  each  side  of  K  fit  the  gear 
wheels  A  and  B. 

Fig.  145  shows  the  diverging  nozzles  of  the  De  Laval  turbine. 
The  function  of  these  nozzles  is  to  reduce  the  pressure  of  the 
steam  by  expanding  it.  At  the  same  time  the  velocity  of  the 
steam  is  increased.  In  other  words,  the  energy  of  pressure 
which  the  steam  contains  before  entering  the  nozzles  is  changed 
in  the  nozzles  into  the  energy  of  velocity.  The  steam  issues 
from  the  nozzles  at  a  very  high  velocity.  It  has  been  shown  that 


HEAT  ENGINES 


FIG.  145. — De  Laval  turbine  wheel. 


L-89 


Driven 

Coupling 

Driving 
Coupling 


146. — Cross-section  of  De  Laval  turbine. 


STEAM  TURBINES 


253 


the  best  efficiency  of  the  turbine  of  this  type  occurs  when  the 
peripheral  speed  of  the  wheel  is  about  half  the  speed  of  the  steam 
leaving  the  nozzles;  The  peripheral  speed  of  the  wheel  must  then 
be  very  high  in  order  to  obtain  good  efficiency. 

In  order  to  bring  the  speed  of  the  turbine  wheel  within  practical 
limits  for  utilizing  the  power,  the  reduction  gears  A  and  B  in 
Fig.  146  are  required.  This  reduction  is 
usually  about  ten  to  one,  and  is  accom- 
plished by  means  of  small  pinions  in  the 
shaft  meshing  with  the  gear  wheels.  The 
teeth  are  cut  spirally,  on  one  side  with  a 
right-hand,  and  on  the  other  with  a  left- 
hand  spiral.  This  method  effectually  pre- 
vents any  movement  of  the  shaft  in  the 
direction  of  the  axis  and  balances  the 
thrust  of  the  gears. 

On  account  of  the  very  high  speeds  at 
which  De  Laval  turbines  operate,  blade 
wheels,  shaft,  and  bearings  require  very 
careful  designing.  The  strength  of  the 
disk,  or  a  wheel  of  a  disk  type,  in  which 
there  is  a  hole  at  the  center,  is  at  best  not 
more  than  half  as  great  as  one  without  a 
hole.*  On  this  account  the  larger  sizes  of 
De  Laval  turbine  wheels  are  made  without 
aj  hole  at  the  center.  The  shaft  is  made 
with  large  flanged  ends  which  are  bolted 
into  suitable  recesses  in  the  hub. 

A  simple  throttling  governor  is  used  for 
speed  regulation  in  the  De  Laval  turbines. 
By  reducing  the  pressure  of  the  steam 
admitted  to  the  nozzles  at  light  loads,  the 
steam  is  discharged  upon  the  blades  at  a 
lower  velocity  than  when  it  is  at  the  higher 
pressure,  and  correspondingly  less  energy  is  given  to  the  turbine 
wheel  so  as  to  maintain  a  constant  speed. 

Fig.  147  shows  the  variation  of  the  velocity  and  the  pressure  in 
the  nozzles  and  blades  of  a  De  Laval  single-stage  turbine.  Curve 
II  shows  the  velocity  for  each  point  in  the  nozzle.  The  ordinates 
represent  the  velocity,  and  the  abscissae  the  position  in  the  nozzle 

*  See  Moyer's  The  Steam  Turbine,  page  333. 


SECTION  A 


K 


SECTION  B 

FIG.  147. — Variation 
of  velocity  and  pres- 
sure in  a  De  Laval  tur- 
bine. 


254 


HEAT  ENGINES 


or  blade.  In  a  similar  manner,  the  change  of  the  pressure  is 
shown  in  Curve  I.  The  figure  shows  that  the  velocity  at  the 
entrance  to  the  nozzle  is  almost  zero  and  the  pressure  a  maximum. 


FIG.  148. — Casing  of  De  Laval  velocity-stage  impulse  turbine. 

When  the  steam  issues  from  the  nozzle,  the  velocity  is  maximum 
and  the  pressure  minimum. 

De  Laval  turbines  are  also  made  of  the  velocity-stage  impulse 
type  in  which  the  steam  is  expanded  from  the  initial  to  the  final 


FIG.  149. — Nozzles  of  Curtis  turbine. 

pressure  in  one  set  of  nozzles,  but  the  velocity  is  absorbed  in  two 
sets  of  moving  blades  with  a  set  of  stationary  ones  between;  and 
in  the  pressure-stage  impulse  or  multicellular  type  where  the 
steam  expands  through  "  successive  sets  of  nozzles  with  corre- 


STEAM  TURBINES 


255 


spending  pressure  steps,  the  velocity  produced  in  each  stage  being 
expended  upon  a  corresponding  row  of  moving  buckets."  This  is 
really  a  "  series  of  single-stage  wheels  each  enclosed  in  a  separate 
cell  or  compartment  and  all  mounted  upon  a  common  shaft." 

These  two  types  of  De  Laval  turbines 
are  similar  respectively  to  a  single-stage 
Curtis  turbine  and  a  Rateau  turbine, 
both  of  which  will  be  described  later  on. 

Fig.  148  shows  the  casing  of  a  velocity- 
stage  turbine  after  the  rotor  has  been 
removed.  The  individual  nozzles  by 
which  steam  is  directed  upon  the  first 
row  of  moving  buckets  and  the  inter- 
mediate stationary  guide  vanes  by  which 
the  steam  is  redirected  upon  the  second 
row  of  moving  buckets  are  seen  in  place. 

173.  The  Curtis  Turbine.— In  the 
Curtis  turbine  as  in  the  De  Laval,  the 
steam  is  expanded  in  nozzles  before 
reaching  the  moving  blades,  but  the 
complete  expansion  from  the  boiler  to 
the  exhaust  pressure  occurs  usually  in  a 
series  of  stages,  or  steps,  as  the  steam 
passes  through  a  succession  of  chambers 
separated  from  each  other  by  dia- 
phragms. In  very  small  sizes  of  the 
Curtis  turbines,  there  is  usually  only 
one  pressure  stage,  but  in  larger  sizes 
there  are  from  two  to  five. 

The  nozzles  of  the  Curtis  turbine  are 
generally  rectangular  in  cross-section, 
and,  because  they  are  always  grouped  close  together,  they  are 
either  cast  integral  with  the  diaphragms  or  in  separate  nozzle 
plates  (Fig.  149),  which  in  assembling  are  bolted  to  the  dia- 
phragms. Most  Curtis  turbines  are  made  with  horizontal 
shafts,  though  the  larger  sizes  often  have  the  shafts  placed  verti- 
cally. In  these  vertical  turbines  the  weight  of  the  turbine  is 
supported  on  a  special  foot-step  bearing  which  carries  the  shaft 
on  a  thin  film  of  oil  supplied  to  the  bearing  under  pressure. 

It  is  typical  of  these  turbines  that  there  are  always  three  or 
more  rows  of  blades,  or  "  buckets,"  following  each  group  of  noz- 


SECTION  8 

FIG.  150. — Variation  of 
velocity  and  pressure  in  a 
single-stage  Curtis  turbine. 


256 


HEAT  ENGINES 


zles,  and  one  of  these  rows  is  stationary.  This  arrangement  in  the 
single-stage  turbine  is  illustrated  in  Fig.  150.  No  expansion 
takes  place  in  the  stationary  blades,  and  the  object  in  using  sev- 
eral rows  of  blades  is  on.ly  to  reduce  the  velocity  to  be  absorbed 
per  row,  and  consequently  to  reduce  the  peripheral  speed  of  the 
wheels  necessary  to  attain  the  best  efficiency. 

Fig.  151  shows  the  path  of  the  steam  through  the  blades 
or  buckets  in  a  Curtis  turbine,  and  Fig.  152  shows  the  plan 
and  elevation  of  a  two-stage  Curtis  turbine. 


FIG.  151. — Path  of  steam  through  moving  and  stationary  buckets 
in  a  Curtis  turbine. 

The  speed  of  the  turbine  is  controlled  by  a  governor  that  "cuts 
out"  the  nozzles,  the  number  of  nozzles  discharging  steam 
through  the  turbine  blades  being  determined  by  the  governor. 
The  Curtis  turbine  is  made  in  a  large  range  of  sizes,  being  sold  in 
sizes  from  15  to  30,000  kw.  The  most  common  application  of 
these  turbines  is  to  the  driving  of  electric  generators. 

A  section  of  a  9000  kw.  Curtis  vertical  turbine  generator  is 
shown  in  Fig.  153.  This  figure  shows  the  electric  generator  at  the 
top  of  the  figure,  the  diaphragms  and  the  wheels  of  the  five  pres- 


STEAM  TURBINES 


257 


sure  stages  immediately  below,  and  at  the  bottom,  the  step- 
bearing  at  the  end  of  the  vertical  shaft. 

174.  The  Rateau  Turbine. — The  Rateau  turbine  has  been 
termed  "  Multicellular,"  that  is,  it  consists  of  a  large  number  of 
cells,  or  pressure  stages,  of  which  each  stage  is  like  a  separate 


ELEVATION    I 


FIG.  152.- 


-Plan  and  elevation  showing  path  of  steam  in  a  two -stage  Curtis 
turbine. 


single-stage  De  Laval  turbine.  Each  stage  contains  one  row  of 
blades,  so  that  the  velocity  that  can  be  absorbed  efficiently  by 
each  stage  is  less  than  in  the  Curtis,  and  the  turbine  must  contain 
a  larger  number  of  stages  between  the  same  limits  of  pressure. 

Fig.  154  shows  diagrammatically  a  Rateau  turbine  with  two 
groups  of  nozzles,  and  therefore  with  two  pressure  stages.     Steam 

17 


258 


HEAT  ENGINES 


FIG.  153. — Section  of  9,000  kilowatt  Curtis  turbine-generator. 


STEAM  TURBINES 


259 


at  the  initial  pressure  enters  the  first  group  of  nozzles  and  expands 
to  the  pressure  of  the  first  stage.  In  this  expansion  it  delivers  a 
portion  of  its  energy  to  the  blades.  It  then  expands  to  the 
exhaust  pressure  in  the  second  group  of  nozzles,  shown  in  the  dia- 
gram between  the  first  and  second  stages. 

In  the  commercial  turbines  of  this  fcypethe  pressure  drop  at  each 
stage  is  small  and  the  nozzles  are  always  made  with  a  uniform 


HV 

SECTION  B 

FIG.  154. — Variation  of  velocity  and  pressure  in  a  two-stage  Rateau 

turbine. 

cross-section  along  their  length;  or  in  other  words,  they  are 
"  non-expanding." 

Fig.  155  shows  four  typical  stages  of  a  Rateau  turbine,  and  Fig. 
156  shows  a  cross-section  of  a  Southwark- Rateau  turbine. 

175.  Kerr  Turbine. — Impulse  turbines  with  bucket  wheels  of 
the  Pelton  type  have  been  developed  to  the  commercial  stage. 
The  Kerr  turbine  was  formerly  of  the  Pelton  type  and  was  the 
most  characteristic.  But  as  constructed  at  the  present  time  it  is 


260 


HEAT  ENGINES 


practically  of  the  Rateau  type.  Fig.  157  shows  a  bucket  wheel  and 
nozzles  of  the  older  type;  and  Fig.  158  shows  the  new  type  of  Kerr 
turbine. 

176.  The  Sturtevant  Turbine. — Another  steam  turbine  in 
which  the  steam  jets  are  discharged  in  a  radial  direction  upon 
the  bucket  wheel  is  known  as  the  Sturtevant  turbine.  In  this 
respect  it  resembles  the  older  Kerr  turbine.  Fig.  159  is  a  good 
illustration  of  this  turbine,  showing  the  buckets  on  the  wheel  and 
the  "reversing"  buckets  on  the  inside  of  the  casing.  These  re- 


FIG.  155. — Four  stages  of  Rateau  turbine. 

versing  buckets  are  not  cut  all  the  way  around  the  circumfer- 
ence, but  three,  four,  or  five  are  cut  following  each  nozzle,  depend- 
ing on  the  velocity  of  the  steam.  The  buckets  are  cut  out  of  the 
solid  metal  of  the  rim  of  the  wheel,  which  is  a  single  forging  of 
open-hearth  steel.  By  this  construction  a  wheel  of  great  strength 
is  secured  and  blade  breakage  is  practically  eliminated.  This 


STEAM  TURBINES 


261 


262 


HEAT  ENGINES 


turbine  was  designed  in  all  its  parts  to  require  the  minimum 
amount  of  attention  and  repairs.  It  is  stated  that  it  can  be  op- 
erated continuously  under  ordinary  conditions  with  little  more 
attention  than  that  required  for  filling  the  oil-wells  once  a 
week. 

177.  The  Parsons  Turbine. — The  Parsons  turbines  are  the 
only  commercial  turbines  of  the  reaction  type,  and  these  operate 
only  partially  on  this  principle.  In  these  turbines  the  stationary 
blades  take  the  place  of  the  nozzles  in  other  forms,  and  direct  the 


FIG.  157. — Bucket  wheel  of  a  Kerr  turbine. 

steam  upon  the  moving  blades.  The  system  of  blading  in  the 
Parsons  turbine  is  shown  in  Fig.  160.  Steam  enters  from  the 
admission  space  as  shown  in  the  figure,  and  passes  through  the 
stationary  blades  where  it  expands  with  an  increased  velocity. 
From  these  blades  it  is  passed  to  the  first  set  of  moving  blades, 
in  which  it  again  expands.  The  variation  of  velocity  and  pressure 
in  passing  through  one  of  these  turbines  is  clearly  shown  in 
Fig.  160. 

A  section  of  one  of  the  simplest  Parsons  turbines  is  shown  in 
Fig.  161.  The  rotating  part  is  a  long  drum  of  three  different  sec- 
tions supported  on  two  bearings — one  at  each  end.  Rows 
of  moving  blades  are  mounted  on  the  circumference  of  this  drum 
and  corresponding  stationary  blades  are  fitted  to  the  inside 
of  the  turbine  casing.  An  annular  space  A  is  a  steam  chest  which 
receives  high-pressure  steam  from  the  steam  mains.  From  this 


STEAM  TURBINES 


263 


annular  space,  the  steam  passes  through  alternate  rows  of  moving 
and  stationary  blades  to  the  exhaust  at  B.  There  are  also  two 
other  annular  spaces  where  the  section  of  the  drum,  or  rotor,  is 
increased  in  diameter,  and  at  these  places  is  an  unbalanced  pres- 
sure, or  thrust  toward  the  right  (in  this  design)  caused  by  the 
pressure  of  the  steam.  This  thrust  is  increased  by  the  expansion 


FIG.  158. — Cross-section  of  new  type  Kerr  turbine. 

of  the  steam  in  unsymmetrical  blades.  To  balance  this  axial 
pressure,  three  balance  pistons  are  provided  at  the  left  end  of  the 
casing — one  for  each  section  of  the  rotor.  Passages  are  cored  out 
in  the  casing  to  make  each  balance-piston  communicate  with  its 
corresponding  section  of  the  rotor,  so  that  the  steam  pressure  on 
each  piston  is  approximately  the  same  as  that  in  the  correspond- 


264 


HEAT  ENGINES 


ing  section.  Except  for  some  differences  in  the  design  of  mechani- 
cal details,  the  turbine  shown  in  Fig.  161  represents  very  well  the 
usual  Parsons  type.  The  Parsons  turbine  is  governed  by  admit- 
ting the  steam  in  puffs.  The  interval  of  time  between  the  puffs 
decreases  as  the  load  in- 
creases, until,  when  the 
overload  capacity  of  the 
turbine  is  reached,  the 
steam  is  admitted  in  a 
practically  continuous 
stream. 

178.  "Impulse  and  Re- 
action" Double-flow  Tur- 
bines.— Recently  a  design 


Drum  Rotor 

•Hi 

SECTION  A 


FIG.  159. — Sturtevant  turbine. 


SECTION  B 

FIG.  160. — Variation  of  velocity  and 
pressure  in  a  Parsons  turbine. 


of  double-flow  turbine  has  been  adopted  by  the  Westing- 
house  Company  for  large  sizes  to  replace  the  single-flow  Par- 
sons type.  There  are  two  principal  advantages  resulting  from 
this  change:  (1)  end  thrust  is  practically  eliminated;  and  (2) 
the  impulse  element  reduces  very  considerably  the  length  of  the 
turbine. 

jFig.  162  illustrates  such  a  double-flow  turbine  with  an  impulse 
element.  In  its  essential  parts  this  turbine  consists  of  a  group  of 
nozzles,  an  impulse  wheel  with  three  rows  of  blades — two 
moving  and  one  stationary, — and  two  intermediate  and  two  low- 


STEAM  TURBINES 


265 


266 


HEAT  ENGINES 


pressure  sections  of  typical  Parsons,  or  " reaction"  blading. 
Steam  is  admitted  to  the  turbine  through  the  nozzle  block  or 
chamber  at  the  bottom  of  the  figure,  and  is  discharged  from  the 
nozzles  at  a  very  high  velocity  to  impinge  on  the  impulse  blades. 
After  passing  through  these  blades,  it  divides,  one  half  expand- 
ing through  the  intermediate  and  low  pressure  Parsons  blading 
at  the  right  of  the  impulse  wheel,  and  the  other  half  expanding 
through  the  blading  to  the  left.  After  leaving  the  low-pressure 
blading  it  exhausts  through  the  passages  E,  E  into  the  condenser. 


FIG.  162. — Double-flow  Westinghouse  turbine. 

179.  Low-pressure  Turbines. — The  high  vacuums  that  can  be 
obtained  with  the  modern  condenser  have  led  to  the  development 
of  the  low-pressure  turbine  in  which  the  pressure  range  is  entirely 
below  atmospheric  pressure. 

"A  pound  of  steam,  expanded  in  a  perfect  heat  motor  from 
ordinary  boiler  pressures  to  a  28-in.  vacuum,  will  develop  about 
one -half  of  the  total  work  in  the  range  of  expansion  above  atmos- 
pheric pressure  and  the  other  half  in  the  range  of  expansion 
from  atmospheric  pressure  to  vacuum.  The  ordinary  recipro- 
cating engine  operates  with  fair  efficiency  in  the  range  above 
atmospheric  pressure,  but  fails  to  develop  more  than  about  one- 
third  of  the  work  theoretically  available  below  atmospheric 
pressure.  This  is  partly  because  of  the  narrowness  of  the  exhaust 
ports  and  the  alternate  cooling  and  heating  of  the  cylinder  walls, 
but  principally  because  of  the  restricted  volume  of  the  low- 
pressure  cylinder,  or  rather,  the  limitations  which  are  placed 
on  the  ratio  of  expansion.  The  steam  turbine,  on  the  other 
hand,  is  not  hampered  in  this  way,  as  it  is  a  simple  matter  to 
provide  all  the  area  required  by  the  steam  at  the  lowest  condenser 
pressure,  and  the  alternate  heating  and  cooling  of  metal  surfaces 


STEAM  TURBINES  267 

are  avoided.  With  most  types  of  turbines  the  efficiency  ratio 
below  atmospheric  pressure  is  better  than  that  obtainable  with 
any  other  type  of  steam  motor  through  any  range  of  pressure. 
It  thus  comes  about  that  low-pressure  steam  turbines  used  in 
conjunction  with  efficient  high-pressure  reciprocating  engines 
at  present  hold  the  published  record  for  highest  efficiency  ratio 
in  large  sizes." 

"In  practice  it  is  found  that  a  low-pressure  turbine  of,  say,  300 
horse-power  capacity,  will  develop  a  horse-power  hour  on  about 
28  Ibs.  of  steam  at  atmospheric  pressure,  exhausting  into  a 
vacuum  of  28  in.  In  other  words,  the  output  of  reciprocating 
engine  plants,  at  present  running  non-condensing,  can  be  in- 
creased 100  per  cent,  through  the  use  of  exhaust  turbines,  without 
requiring  the  generation  of  more  steam  or  the  burning  of  more 
fuel.  Where  simple  engines  are  at  present  running  condensing, 
the  output  of  power  per  pound  of  steam  can  be  increased  by 
60  per  cent,  and  the  output  of  compound  condensing  engines  by 
a  somewhat  less  amount — about  25  per  cent.  By  lengthening 
the  period  of  admission  of  the  engines,  the  power-producing 
capacity  can  be  increased  in  a  still  greater  ratio." 

180.  Mixed  Flow  or  Mixed  Pressure  Turbines. — "Turbines 
in  which  low  pressure  steam  is  admitted  to  an  intermediate 
stage  from  some  exterior  source,   as  the  exhaust  of  a  steam 
engine,  are  known  as  mixed  flow  or  mixed  pressure  turbines. 
They  are  usually  designed  to  operate  normally  with  low  pressure 
steam,  admitting  high  pressure  steam  only  as  required  in  case 
of  a  deficiency  of  the  low  pressure  steam  supply.     They  differ 
from   standard   high   pressure   condensing  turbines   principally 
in  the  different  proportioning  of  the  areas  through  the  two  parts 
of  the  turbines  and  in  the  provision  of  a  special  governing  gear 
for  automatically  controlling  the  admission  of  the  live  steam. 
In  case  of  complete  failure  of  the  low  pressure  steam  supply, 
the  turbine  will  operate  with  good  economy  on  high  pressure 
steam  alone.     Other  arrangements  are  also  supplied  to  meet 
special   conditions." 

181.  Bleeder  Turbines.— Turbines  from  which  steam  is  drawn 
from  an  intermediate  stage  for  use  in  heating,  raising  the  tem- 
perature of  the  feed  water,  etc.,  are  called  bleeder  turbines. 

182.  Application  of  the  Steam  Turbine. — The  steam  turbine 
is  adapted  primarily  to  the  driving  of  machines  which  require  a 
high  rotative  speed.     They  are  not  applicable  where  a  large 


268  HEAT  ENGINES 

starting  effort  is  required,  as  the  forces  acting  in  the  turbine 
are  relatively  small.  Their  use  is  therefore  principally  in  driving 
machines  which  start  without  load,  such  as  electric  generators, 
centrifugal  fans,  and  propeller  wheels.  The  turbine  is  not 
suited  to  the  propelling  of  vehicles,  or  to  the  driving  of  mills  by 
belt  drive.  Such  applications  of  power  require  more  initial 
starting  effort  than  the  turbine  is  capable  of  producing.  The 
steam  turbine  is  also  coming  into  extensive  use  in  marine  work. 
Its  uniform  torque  and  freedom  from  vibration  are  points  which 
make  it  highly  suitable  to  this  field. 


CHAPTER  XIV 
THE  INTERNAL  COMBUSTION  ENGINE 

183.  Historical. — The  internal  combustion  engine  has  now  been 
in  commercial  use  for  over  fifty  years.  During  this  time  many 
improvements  have  been  made  in  its  certainty  of  operation,  regu- 
lation and  fuel  economy.  In  the  earlier  engines  the  gas  consump- 
tion was  about  100  cu.  ft.  per  horse-power  for  each  brake  horse- 
power developed;  to-day  the  consumption  of  gas  in  the  most 
economical  engines  has  been  reduced  to  as  low  as  14  cu.  ft.  per 
brake  horse-power  per  hour.  This  may  be  expressed  in  another 
way,  that  is,  the  heat  efficiency  of  the  internal  combustion  engine 
has  been  increased  in  fifty  years  from  4  per  cent,  to  30  per  cent. 
The  regulation  has  been  slowly  improved  so  that  now  it  is  possi- 
ble to  operate  gas  engines  for  service  requiring  most  exacting 
speed  control  and  its  certainty  of  operation  is  almost  as  satisfac- 
tory as  that  of  the  steam  engine. 

During  the  first  twenty  years  of  the  development  of  internal 
combustion  engines  most  of  the  engines  built  were  small  in  size, 
not  exceeding  10  horse-power;  but  to-day  this  type  of  engine  is 
being  built  in  sizes  of  over  5000  brake  horse-power.  There  are 
in  use  in  the  world  to-day  over  800,000  internal  combustion  engine 
installations.  The  history  of  the  development  of  the  gas  engine 
is  extremely  interesting  and  profitable  reading  and  is  well  treated 
in  the  works  of  Dugald  Clerk,  Hugo  Giildner  and  others. 

The  first  important  commercial  gas  engines  were  placed  on 
the  market  by  Otto  in  1878  and  the  Otto  engine  was  the  first 
engine  in  which  the  gases  were  compressed  in  the  cylinder  pre- 
vious to  compression,  and  this  type  of  engine  has  been  the  pre- 
vailing type  in  use.  In  1898  an  engine  built  in  accordance  with 
the  principles  laid  down  by  Rudolf  Diesel  and  using  fuel  oil 
showed  a  much  higher  efficiency  than  had  been  obtained  hereto- 
fore. This  engine  has  opened  up  a  still  wider  field  for  the  use 
of  the  gas  engine  and  is  now  being  extensively  used  in  stationary 
practice  and  is  being  introduced  in  marine  practice.  The.  internal 
combustion  engine  is  also  being  used  in  connection  with  the  pro- 

269 


270  HEAT  ENGINES 

duction  of  power  from  the  waste  gases  of  blast  furnaces  and 
coke  ovens.  There  are  still  many  other  applications  of  internal 
combustion  engines  which  will,  in  the  future,  increase  its  already 
extensive  use  as  a  prime  mover. 

184.  Classification   of   Gas   Engines. — Gas   engines   may   be 
divided  into  two  general  types; 

(1)  those  in  which  the  ignition  occurs  at  constant  volume; 

(2)  those  in  which  the  ignition  occurs  at  constant  pressure. 
Theoretically  it  is  possible  to  conceive  of  an  engine  working  so 
that  ignition  would  occur  at  constant  temperature,  but  practically 
an  engine  working  in  such  a  cycle  has  not  been  constructed. 
Engines  of  the  first  type  may  be  sub-divided  into  two  classes; 

(a)  those  in  which  the  charge  is  ignited  without  previous 

compression,  and 

(6)  those  in  which  the  charge  is  ignited  with  previous  com- 
pression. 

The  first  internal  combustion  engine  built  belonged  to  type  (1), 
class  (a)  of  which  the  Lenoir  was  a  good  example;  but  owing  to 
its  low  efficiency,  this  class  of  engine  is  no  longer  used. 

185.  Type  (1),  Class  (a). — In  an  engine  of  this  class,  the  charge 
of  gas  and  air  is  drawn  in  at  atmospheric  pressure  for  a  part  of 
its  stroke,  then  the  valve  is  closed  and  the  charge  ignited.     The 
pressure  rises  rapidly,  forcing  the  piston  forward  for  the  remainder 
of  the  stroke.     On  the  return  stroke,  the  products  of  ignition 
are  forced  out  of  the  cylinder.     The  working  cycle  consists  of: 
(1)  feeding  the  cylinder  with  explosive  mixture;  (2)  igniting  the 
charge  of  gas  and  air;  (3)  expanding  the  gases  after  explosion; 
(4)  expelling  the  burned  gases.     Fig.   163  shows  an  indicator 
card  taken  from  an  engine  of  this  class.     Line  AB  is  the  atmos- 
pheric line.     From  A  to  C  the  charge  is  drawn  in,  from  C  to  D 
explosion  occurs,  from  D  to  E  the  gases  expand,  and  along  the 
lines  EB  and  BA   the  gases  are  expelled  from  the  cylinder. 
Owing  to  the  lack  of  previous  compression,  the  pressure  at  D 
must  always  be  comparatively  low,  seldom  exceeding  40  Ibs. 
This,  of  course,  means  lower  initial  temperature  at  the  maximum 
point  of  explosion  and  correspondingly  reduced  economy.     This 
type  of  engine  uses  about  60,000  B.T.U.  per  horse-power  per 
hour. 

A  very  interesting  modification  of  this  type  of  engine  is  the 
free  piston  engine  of  Otto  and  Langen,  which  was  first  shown  at 
the  Paris  Exposition  in  1867.  The  idea  was  not  original  with 


THE  INTERNAL  COMBUSTION  ENGINE        271 

them  but  the  engine  was  an  adaptation  of  an  engine  originally 
proposed  (but  never  built)  by  Barsanti  and  Matteucci  in  Italy 
in  1854.  In  this  particular  type  of  engine  the  cylinder  is  set  verti- 
cally and  the  piston  is  shot  up  by  the  explosion  of  the  gases, 
without  resistance.  On  its  return  stroke  it  engages,  through 
suitable  mechanism,  with  the  fly-wheel  which  is  rotated  by  the 
forces  of  gravity  and  atmospheric  pressure  acting  against  the 
piston  on  this  downward  or  working  stroke. 


FIG.  163.— Type  (1),  class  (a). 

186.  Type  (1),  Class  (b). — In  this  engine,  explosion  occurs 
at  a  constant  volume  with  previous  compression.  This  cycle 
was  first  proposed  by  Beau  de  Rochas  in  1862,  and  was  first 
put  into  successful  operation  by  Otto  in  1876.  In  the  Otto 
engine  a  charge  of  gas  and  air  is  drawn  in  with  the  first  out 
stroke  of  the  engine,  and  on  the  return  stroke  this  charge  is 
compressed.  Near  the  end  of  the  compression  stroke,  the 


FIG.  164.— Type  (1),  class  (6)— Otto  cycle. 

charge  is  ignited,  and  on  the  following  out  stroke,  expands. 
On  the  next  return  stroke  the  charge  is  expelled  from  the  cylinder. 
There  are  five  operations  in  this  cycle:  (1)  charging  the  cylinder 
with  gas  and  air;  (2)  compressing  the  charge  in  the  clearance 
space;  (3)  igniting  the  mixture;  (4)  expanding  the  hot  gases  after 
ignition;  (5)  expelling  the  burned  gases  on  the  exhaust  stroke. 
A  diagram  of  this  cycle  is  shown  in  Fig.  164.  Along  the  line 
AB  the  charge  is  drawn  in;  along  the  line  BC  it  is  compressed; 


272  HEAT  ENGINES 

along  CD  it  is  ignited;  along  DE  it  is  expanded;  and  along  EA 
it  is  expelled  from  the  cylinder. 

Otto  compressed  the  gases  previous  to  ignition  but  did  not 
realize  that  this  was  in  itself  the  reason  for  the  marked  economy 
which  his  engine  showed  over  all  other  engines  that  had  previously 
been  built.  He  attributed  the  economy  of  his  engine  to  the  com- 
pression of  the  gas  producing  a  thorough  mixture  and  preventing 
the  gas  and  air  from  being  in  layers  in  the  cylinder. 

The  cycle  just  described  is  that  of  &  four-cycle  engine,  that  is, 
to  complete  the  cycle  of  the  gases  requires  four  strokes  of  the 
engine.  In  the  two-cycle  engine  the  same  cycle  of  events  for 
the  gases  is  completed  in  two  strokes. 

The  method  of  obtaining  the  full  cycle  of  events  in  two  strokes 
in  the  smaller  types  of  engines  is  illustrated  in  Fig.  165. 


1.  2. 

FIG.  165. — Cross-section  of  two-cycle  engine. 

Figure  165-1  shows  the  engine  in  its  initial  position.  The 
fly-wheel  has  previously  been  rotated,  drawing  a  charge  of  gas 
and  air  through  port  I  into  the  engine  base  and  partly  com- 
pressing it  there.  In  the  larger  engines  this  charge  of  gas  and 
air  is  externally  compressed.  In  the  position  shown  the  fuel 
charge  previously  compressed  in  the  base  is  entering  the  cylinder 
through  the  port  1%.  At  the  same  time  the  burned  gases  are  leav- 
ing the  cylinder  through  the  exhaust  port  Is.  A  deflecting  plate 
P  on  the  head  of  the  cylinder  prevents  the  incoming  gases  from 
passing  out  of  the  exhaust  port.  As  the  piston  moves  up  it 
covers  ports  Iz  and  Is,  and  further  compresses  the  gas  into  the 
clearance  space.  As  the  piston  approaches  the  upper  dead 
center,  shown  in  Fig.  165-2,  the  gases  are  ignited.  This  ignition 


THE  INTERNAL  COMBUSTION  ENGINE        273 

occurs  when  the  piston  is  practically  stationary  so  that  it  may 
be  said  to  occur  at  constant  volume.  The  pressure  produced  by 
the  rapid  heating  of  the  gases  due  to  ignition,  forces  the  piston 
down  as  soon  as  the  dead  center  is  passed.  While  the  piston 
is  in  the  position  shown  in  Fig.  165-2,  port  /  is  fully  uncovered 
and  the  vacuum  produced  by  the  upward  movement  of  the  piston 
draws  air  and  gas  through  the  carbureter  into  the  crank  case 
so  as  to  furnish  fresh  gas  and  air  for  the  next  cycle  of  the  engine. 
During  the  following  downward  stroke  of  the  piston  the  burned 
gases  above  the  piston  are  allowed  to  expand,  and  the  fresh 
charge  below  it  is  compressed.  This  stroke  brings  the  piston 
to  the  position  shown  in  Fig.  165-1  and  the  cycle  of  operations 
is  thus  completed  in  two  strokes  instead  of  four,  as  described 
for  a  four-cycle  engine. 

187.  Type   (2).  —  The  most  representative  engine  of  this  type 
in  commercial  use  is  the  Diesel. 
A  typical  indicator  card  of  this 
engine  is  shown  in  Fig.  166. 

In  this  engine  air  is  drawn  in  on 
the  downward  stroke  and  then 
compressed  on  the  next  up  stroke 
to  a  pressure  of  about  500  Ibs. 
per  square  inch.  The  compres- 
sion  of  the  gas  to  this  high  pres- 


sure  increases  its  temperature  to     FlG'  m-  Card  °f 


over  1000°  F.  At  the  beginning 
of  the  next  down  stroke  of  the  engine  fuel  oil  previously  com- 
pressed to  a  pressure  of  about  800  Ibs.  is  injected  into  the  cylinder 
and  owing  to  the  high  temperature  of  the  air  in  the  cylinder,  the 
oil  on  entering  is  ignited.  The  heat  produced  by  this  ignition  of 
the  oil  maintains  the  pressure  in  the  cylinder  up  to  the  point  at 
which  the  oil  supply  is  cut  off  by  the  governor.  For  the  remain- 
der of  this  stroke  the  gases  in  the  cylinder  are  expanded.  On  the 
fourth  stroke  of  the  engine  the  gases  are  exhausted  from  the  cyl- 
inder. The  cycle  of  operation  is  shown  in  Fig.  166  in  which  the 
line  AB  represents  admission,  BC  compression,  CD  the  period  of 
fuel  admission  controlled  by  the  governor,  DE  expansion,  and  EA 
represents  exhaust.  The  cycle  just  described  is  for  a  four-cycle 
motor,  but  this  same  cycle  may  be  used  in  a  two-cycle  motor. 
In  the  two-cycle  motor  during  the  first  stroke  the  air  will  be  com- 
pressed, on  the  next  stroke  expansion]  (for  about  75  per  cent. 

18 


274 


HEAT  ENGINES 


L_ 

c 

^T 

rt 

r 

1 

I 

III 

r 

DO 

DJIJDDDDDC: 

c 

3D 

D 

(II  1 

ii  M  i   ' 

ffltftffmn 

THE  INTERNAL  COMBUSTION  ENGINE        275 

of  the  stroke),  exhaust  and  admission  will  occur.  For  the 
two-cycle  Diesel  engine  the  air  entering  the  cylinder  is  com- 
pressed by  the  external  pump  to  about  8  Ibs.  and  this  air  assists  in 
driving  out  the  burned  gases  or,  as  it  is  called,  scavenging  the 
cylinder. 

Until  recently  the  Diesel  engines  have  been  vertical,  but  now 
horizontal  engines  are  also  being  made  of  this  type. 

Fig.  167*  shows  a  four-cylinder  Diesel  engine  for  marine  pur- 
poses. A  vertical  air  pump  is  coupled  to  one  end  of  the  crank 
shaft  and  compresses  air  in  two  stages  from  750  to  100Q  Ibs. 
pressure.  A  governor  is  provided  that  adjusts  the  speed  from  150 
to  400  r.p.m.  By  reducing  the  quantity  of  fuel  delivered  by  the 
pumps  the  speed  can  be  still  further  reduced.  The  working  cycle 
and  details  of  operation  are  essentially  the  same  as  in  the  older 
types  of  stationary  Diesel  engine.  The  engine  shown  in  the  figure 
is  a  recent  type  of  high-speed  Diesel  engine  and  develops  about 
250  H.  P.,  is  11  ft.  10  in.  long,  39  in.  wide  and  7  ft.  high.  The 
weight  of  the  engine  is  about  26,000  Ibs.  The  actual  test  of 
this  engine  showed  a  consumption  of  about  J  Ib.  of  fuel  oil  per 
brake  horse-power  per  hour  at  350  r.p.m.  and  a  full  load.  The 
mechanical  efficiency  at  full  load  was  about  78  per  cent.  This 
engine  converted  about  34  per  cent,  of  the  total  heat  of  the  fuel 
into  available  mechanical  energy.  Its  theoretical  efficiency  was 
56  per  cent.,  so  that  about  60  per  cent,  of  the  theoretically 
available  heat  was  actually  utilized  in  the  engine. 

188.  Theoretical  Cycles. — In  determining  the  mathematical 
expression  for  efficiency  from  the  theoretical  cycles  for  the  two 
principal  types  of  engine  we  must  in  each  case  distinguish  two 
conditions;  one  in  which  the  gas  is  expanded  completely  to  the 
exhaust  pressure  and  one  in  which  the  gas  is  partially  expanded 
and  exhaust  begins  at  a  pressure  higher  than  the  exhaust 
pressure  line. 

Fig.  168  represents  an  engine  of  type  (1),  class  (b)  working 
with  complete  expansion.  Lines  DE  and  CB  are  assumed  to  be 
adiabatics,  line  CD  is  a  constant  volume  line,  line  BE  is  a  con- 
stant pressure  line.  All  the  heat  will  be  absorbed  along  the  line 
CD  and  all  the  heat  will  be  rejected  along  the  line  EB.  Let  the 
heat  received  along  line  CD  be  represented  by  H  i  and  the  abso- 
lute temperature  at  C  by  Tc  and  at  D  by  Td;  let  the  heat  rejected 

*  Figure  taken  from  Power,  August  15,  1911,  p.  248. 


276 


HEAT  ENGINES 


along  EB  be  represented  by  Hz  and  the  absolute  temperature 
at  E  by  Te  and  at  B  by  Tb;  and  let  the  weight  of  the  charge  =  w. 

Then  Hl  =  wcv(Td  -  Tc) 

and  Hz  =  wcp(Te  -  Tb) 

The  work  done  =  HI  -  Hz  =  wcv(Td  -  Tc)  -  wcp(Te  -  Tb) 

)   -  wcp(Te  -  Tb) 


.               Hi  -  H2      wcv(Td  — 
Efficiency  = ff~~ 


=  1-7 


Te-Tb 


wcv(Td  -  Tc) 


(1) 


Td-  Tc 

From  Fig.  168  it  will  be  seen  that  for  any  compression  temper- 
ature Tc  and  given  rise  in  temperature  during  ignition  there  is  a 

Y 


FIG.  168. — Type  (1),  class  (6) — with  complete  expansion. 

temperature  Te  at  which  the  adiabatic  line  touches  the  back 
pressure  line  AB.  We  can,  therefore,  establish  a  relation  be- 
tween Tb,  TC}  Td  and  Te.  If  we  allow  the  pressure  and  volume 
at  each  particular  p  oint  to  be  represented  by  P  and  V  respec- 
tively, with  a  subscript  corresponding  to  the  letter  at  that  point, 
then 

PdVJ  =  PeVy  (2) 

and  PcVcy  =  PbVb.y  (3) 

Dividing  (2)  by  (3) 


But 
therefore 

and  since 

then 

and 


Vd  =  Vc  and  Pe  =  Pb, 
Pd      V7 


Pd  =  Tc 

T±   _    Te 

~Tc~~Tt 

Te   =    Tl 


Ve  Te 

Fi         K 


(4) 


THE  INTERNAL  COMBUSTION  ENGINE        277 


Although  this  represents  the  best  cycle  of  type  (1),  class  (b) 
engine,  it  is  not  one  commonly  used  and  as  yet  no  engine  has  been 
put  upon  the  market  using  this  cycle. 

Fig.  169  represents  the  ideal  diagram  of  the  Otto  engine. 
In  this  diagram  the  expansion  is  only  carried  far  enough  so  that 
exhaust  commences  when  the  volume  in  the  cylinder  is  the  same 
as  that  before  compression. 

The  lines  BC  and  DE  are  assumed  to  be  adiabatics.  All  the 
heat  must  then  be  absorbed  along  the  line  CD  and  all  rejected 
along  EB.  Let  the  heat  received  along  CD  be  represented  by 


FIG.  169.— Theoretical  Otto  cycle. 

Hi,  and  the  absolute  temperature  at  C  by  Tc,  and  at  D  by  Td; 
let  the  heat  rejected  along  EB  be  represented  by  H%,  and  the 
absolute  temperature  at  E  by  Te  and  at  B  by  Tb'}  and  let  the 
weight  of  the  charge  =  w. 

Then  #1  =  wcv(Td  -  Tc) 
and  #2  =  wcv(Te  -  Tb). 

The  work  done  =H1  -  H2  =  wcv(Td  -  Tc)  -  wcv(Te  -  Tb). 


HI  —  HZ 
Efficiency  =  — ^j— 


wcv(Td-Tc)-wcv(Te-  Tb) 


wcv(Td  -  Tc) 
(Td  -  Tc)  -  (Te  -  Tb) 

Td-fe 
Te-  Tb 


Td-  Tc 

Both  curves  are  adiabatic,  hence 


(5) 


T 


V 


Vc~l 


T 


278  HEAT  ENGINES 

Therefore 

fTf  FT! 

Yd  =  YcJ 

and  by  subtraction, 

rrj  FT?  rrj  fji 

1  e    _     1  e   —    1  b  _    J^b  x«v 

J-  d  J-  d   —    J-  c  J-  c 

Substituting  equation  (6)  in  equation  (5), 

Efficiency  =  1  -  ~-  (7) 

J-  c 

This  is  the  most  important  expression  in  connection  with  the 
gas  engine.  It  shows  that  the  efficiency  of  a  gas  engine  working 
in  the  Otto  cycle  depends  upon  the  temperature  before  and  after 
compression.  The  knowledge  of  this  fact,  first  demonstrated 
by  Dougal  Clerk,  has  led  to  the  production  of  the  modern  high- 
efficiency  engine.  The  same  fact  will  also  be  proved  for  the 
other  types  of  engines. 

The  efficiency  of  the  Otto  cycle  may  also  be  expressed  in  terms 
of  volume  or  pressure. 
Since  the  compression  curve  is  an  adiabatic, 

Tb_  i 

Tc  =:  \Fi 
rri 

Substituting  this  value  of  „  in  equation  (7),  we  have 

•L  c 


. 
Efficiency  =  1  -  (j^)  (8) 


where  r  =  TT  =  the  ratio  of  compression. 

Equation  (8)  shows  that  the  efficiency  depends  upon  and  varies 
with  Vit  the  clearance  volume. 
Finally  since  BC  is  an  adiabatic, 

zv 

Tc  ~ 
Substituting  this  value  in  equation  (7),  we  have 

/Ph\  y^l 
Efficiency  =  1  -  (-—•)    y  (10) 

From  equation  (10)  it  is  seen  that  the  efficiency  increases  as  the 
compression  pressure,  P&,  increases,  since  the  pressure  Pc  will 
remain  nearly  constant. 


THE  INTERNAL  COMBUSTION  ENGINE        279 


In  developing  the  expression  for  the  efficiency  of  an  engine 
working  in  the  Type  (2)  cycle,  it  is  best  to  assume  that  the  com- 
pression takes  place  in  the  same  cylinder  as  the  expansion  and 
exhaust,  although  in  the  actual  engine  two  cylinders  are  used. 
Fig.  170  shows  the  theoretical  diagram  under  this  assumption, 
when  there  is  complete  expansion.  The  lines  BC  and  DE  are 
assumed  to  be  adiabatics.  All  the  heat  must  then  be  absorbed 
along  the  line  CD  and  all  rejected  along  the  line  EB.  Let  the 
heat  received  along  CD  be  represented  by  H\t  and  the  absolute 
temperature  at  C  by  Tc,  and  at  D  by  Td;  let  the  heat  rejected 
along  EB  be  represented  by  Hz,  and  the  absolute  temperature 
at  E  by  Te,  and  at  B  byTb;  and  let  the  weight  of  the  charge  =  w. 


FIG.  170. — Type  (2)  — with  complete  expansion. 

Then  Hl  =  wcp(Td  -  Tc) 

and  #2  =  wcp(Te  -  Tb) 

The  work  done  =  HI  -  H2  =  wcp(Td-  Tc)  --  wcp(Te  -•   Tb) 
1-  #2      wcp  (Td  -  Tc)  -  wcp  (Te  -  Tb) 


Efficiency 


Hl  wcp  (Td  -  Tc) 

(Td  -  Tc)  -  (Te  -  Tb) 

Td-Tc 
Tc-  Tb 


Td- 


(ID 


Both  curves  are  adiabatic,  hence 

Te 

Td 
Therefore 


/ 
/Pe\^  (Pb\ 

=  (pj  y  =  IP;) 


Tb 


and  by  subtraction, 


Te 

Td 


Te-    Tb 

Td  -  Te 


(12) 


280 


HEAT  ENGINES 


Substituting  equation  (12)  in  equation  (11), 

/TT 

Efficiency  =  1  —  --,--.    (see  equation  7) 

1  c 

Since  the  compression  curve  is  an  adiabatic 
Tb  _    /FeV-1       _J_ 

Tc  =:  \vj         =  r^-1 
Tb       /Pb\r=i       /PA  i^-1 


(13) 


Substituting  these  values  for  ^  in  equation  (13),  we  have 

fVc\y~1 

Efficiency  =  1  —  (y~)         (see  equation  8)  (14) 

=  1 ^— [    (see  equation  9)  (15) 

/PA  T-1 
=  1  —  ( p-j   T      (see  equation  10)  (16) 


Y      c 


FIG.  171.— Theoretical  Diesel  cycle. 


In  the  Diesel  engine  as  now  built  ignition  occurs  at 
constant  pressure  and  the  heat  is  rejected  at  constant  volume. 
The  cycle  is  approximately  that  shown  in  Fig.  171.  The  lines 
BC  and  DE  are  assumed  to  be  adiabatics.  All  the  heat  must 
then  be  absorbed  along  the  line  CD  and  all  rejected  along  EB. 
Let  the  heat  received  along  CD  be  represented  by  Hi,  and  the 
absolute  temperature  at  C  by  Tc  and  at  D  by  Td',  and  let  the  heat 
rejected  along  EB  be  represented  by  H2,  and  the  absolute  tem- 
perature at  E  by  Te  and  at  B  by  Tb.  Assume  the  weight  of  the 
charge  to  be  constant  for  the  cycle  and  equal  to  w.  This  assump- 
tion is  permissible  since  the  fuel  is  oil  and  the  increase  in  weight 
of  the  charge  is  small. 
Then  Hi  =  wcp(Td  -Tc) 

and  H2  =  wcv(Te  -  Tb). 


THE  INTERNAL  COMBUSTION  ENGINE        281 


The  work  done  =  Hl  -  Hz  =  wcp(Td  -  Tc)  -  wcv(Te  -  Tb). 

!  -  #2       wcp  (Td  -  Tc)  -  wcv(Te  -  Tb)_ 
^-  WCp  (Td  _  r;} 


Efficiency  = 


-  Te)  -  cv  (T.  -  Tb) 

cp  (Td  -  Tc) 
Te  -  Tb 


y(Td-Tc)' 

Since  CD  is  a  constant  pressure  line  and  EB  a  constant  volume 
line 

Td  =  TcVydc  (18) 

Te       Pe 


and     Yb- 
But     P.  = 


(19) 
(20) 


Substituting  equations  (20)  and  (21)  for  Pe  and  Pb  in  equation 
(19)  we  have 

r.    'I©'      " 

Tb~ 


(ay 
\VJ 


"-**\T.l 

Substituting  equations  (18)  and  (22)  for  Td  and  Te  in  equation 
(17)  we  have 


Efficiency  =  1  — 


=  1- 


1- 


(m   V*          T 

\lcve~ 

T    i  (Vd\y 

Tb  (  (vj 


(23) 


(compare 

,     with  equations  (24) 
9  and  15) 


since 


282 


HEAT  ENGINES 


From  equation  (24)  it  is  seen  that  the  efficiency  of  the  Diesel 
cycle  depends  not  only  upon  the  ratio  of  compression,  r,  but  also 

upon  the  ratio  ^ ,  or  the  ratio  of  the  volume  at  cut-off  to  the 
v c 

clearance  volume. 

189.  Losses  in  the  Gas  Engine. — Fig.  172  shows  the  theoretical 
card  of  a  gas  engine  in  full  lines,  and  the  actual  card  in  dotted 
lines.  The  difference  between  the  actual  and  theoretical  card  is 
largely  due  to  the  losses.  The  actual  compression  line  BF 
differs  from  the  theoretical  line  BC  because  of  the  loss  of  heat  to 
the  cylinder  walls  during  compression.  The  theoretical  line  CD 
assumes  the  combustion  of  the  gas  to  take  place  instantly,  while 
in  actual  operation,  as  shown  by  the  line  FG,  the  burning  of  the 
gas  takes  an  appreciable  length  of  time,  and  may  continue  to  mid- 


': 


FIG.  172. — Difference  between  actual  and  theoretical  indicator  cards. 

stroke.  Due  to  this  fact,  it  is  impossible  to  obtain  full  theoretical 
pressure  at  the  beginning  of  the  stroke.  During  this  operation 
there  is  also  a  loss  of  heat  to  the  walls.  The  expansion  line  DE  in 
the  theoretical  card  is  assumed  to  be  an  adiabatic.  The  actual 
line  GH  is  not  an  adiabatic,  as  after-burning  always  occurs  along 
this  line,  and  in  addition  there  is  a  large  loss  of  heat  to  the  water- 
jacket  surrounding  the  cylinder  of  the  engine. 

There  are  other  losses  in  the  gas  engine  which  are  not  so  appar- 
ent from  the  indicator  card. 

(a)  The  largest  of  all  losses  is  the  loss  of  heat  in  the  exhaust 


THE  INTERNAL  COMBUSTION  ENGINE        283 

gases,  which  leave  the  engine  at  a  high  temperature,  usually  over 
500°. 

(b)  The  next  largest  loss  is  the  heat  carried  away  by  the  water- 
jacket.     This  water-jacket  is  necessary  in  all  stationary  engines 
to  prevent  overheating  of  the  cylinder.     A  similar  loss  occurs  in 
all  air-cooled  cylinders. 

(c)  The  loss  due  to  the  charge  of  gas  and  air  entering  the 
cylinder  being  heated  by  coming  into  contact  with  the  hot  parts  of 
the  engine.     This  heating  of  irhe  charge  increases  its  volume  and 
the  engine  receives  less  weight  of  gas  and  gives  a  correspondingly 
reduced  horse-power. 

(d)  There  is  a  loss  of  effective  pressure  in  the  working  medium 
due  to  the  resistance  in  inlet  and  exhaust  valves. 

The  following  is  a  statement  of  the  distribution  of  heat  in  a  gas 
engine  taken  from  actual  tests  and  expressed  in  per  cent. 

Heat  used  in  indicated  work 25  per  cent. 

Heat  lost  in  exhaust 37  per  cent. 

Heat  lost  in  jacket  water 33  per  cent. 

Heat  lost  in  radiation  and  conduction 5  per  cent. 

The  relative  loss  from  the  exhaust  and  in  the  jacket  varies 
widely  in  different  e,  gines.  In  some  engines  the  exhaust  and 
jacket  losses  are  nearl^  the  same  amount,  and  in  some  the  jacket 
loss  is  even  higher  than  the  loss  in  the  exhaust.  The  loss  in  the 
jacket  may  be  appreciably  decreased  and  the  efficiency  increased 
by  running  the  jacket  water  as  warm  as  successful  operation  will 
permit. 

190.  Gas-engine  Fuels. — The  fuel  used  by  gas  engines  may  be 
classified  under  three  different  heads: 

1.  Solid  fuels. 

2.  Liquid  fuels. 

3.  Gaseous  fuels. 

The  fuel,  no  matter  what  its  original  state  may  be,  must  be 
changed  to  a  gaseous  form  before  it  can  be  used  in  an  engine. 
With  the  first  two  forms  of  fuel,  it  is  necessary  that  some  means  be 
provided  for  vaporizing  them  before  they  are  used  in  the  engine. 

In  the  solid  fuels,  they  are  vaporized  in  some  form  of  gas  pro- 
ducer. They  are  then  used  in  the  engine  as  producer  gas.  In 
the  liquid  fuels,  vaporization  takes  place  in  some  form  of  carbu- 
reter, or  vaporizer,  or  in  the  cylinder  itself. 


284 


HEAT  ENGINES 


191.  Gas  Producers.— In  the  gas  producer,  the  heat  of  the  fuel 
bed  distils  the  volatile  gases  from  the  fresh  coal,  leaving  coke. 
This  coke  is  burned  to  CO  by  introducing  insufficient  air.  A 
small  portion  of  the  carbon  is  changed  to  CC>2.  Producers  using 
anthracite  coal  have  been  in  successful  use  for  a  number  of  years, 
and  bituminous  producers  are  now  coming  into  use.  The  prin- 
cipal difficulty  in  using  bituminous  coal  as  a  fuel  for  producers  is 
in  removing,  or  preventing,  the  formation  of  tar.  The  future 
success  of  the  bituminous  producer  depends  upon  the  thorough 
removal  of  the  tar. 

There  are  two  types  of  producers:  (a)  pressure,  and  (b)  suction 
producers.  In  the  pressure  type,  the  air  and  steam  are  furnished 


FIG.  173. — Cross-section  of  suction  gas  producer. 

to  the  producer  by  a  fan.  The  rate  of  production  is  independent 
of  the  engine's  demand  and  the  gas  must  be  stored.  The  gas  is 
furnished  to  the  engine  at  the  pressure  produced  by  the  fan,  usu- 
ally equivalent  to  a  pressure  of  a  2-  or  3-in.  column  of  water. 

In  the  suction  type,  the  air  is  drawn  through  the  producer  by 
the  suction  formed  in  the  engine  cylinder,  so  that  the  rate  of  pro- 
duction of  gas  in  the  producer  depends  upon  the  demand  of  the 
engine.  The  producer  then  automatically  furnishes  the  necessary 
amount  of  gas  for  the  operation  of  the  engine,  so  that  no  storage 
tank  is  required. 

The  suction  producer  is  becoming  very  popular  for  use  with  the 
gas  engine,  particularly  in  the  smaller  sizes.  The  pressure  pro- 


THE  INTERNAL  COMBUSTION  ENGINE        285 

ducer  is  more  expensive  in  installation  than  the  suction  type,  as 
it  involves  a  gas  holder,  but  it  can  be  used  with  inferior  grades  of 
fuel.  The  suction  producer  occupies  less  space  and  costs  less  than 
the  pressure  type.  It  is  best  adapted  to  the  use  of  high-grade 
fuels.  The  most  successful  suction  producers  use  anthracite  coal. 

Fig.  173  shows  a  cross-section  of  a  suction  producer.  A  is  a 
blower  which  is  used  to  furnish  draft  during  the  starting  of  the 
fires.  B  is  the  generator  with  a  double-valved  hopper  for  admit- 
ting the  coal  to  the  fuel  bed  of  the  producer.  C  is  a  vaporizer  in 
which  steam  is  formed,  the  steam  being  mixed  with  the  air  enter- 
ing the  producer.  D  is  the  scrubber,  consisting  of  a  coke  tower 
with  a  spray  of  water  for  washing  the  gas.  E  is  the  cleaner  con- 
taining trays  filled  with  wood  shaving,  through  which  the  gas 
passes  to  remove  dust  and  dirt.  F  is  the  cleaning  pot  which  col- 
lects the  heaviest  dust  and  dirt  coming  over  with  the  gas.  By 
the  admission  of  water  to  the  cleaning  pot  on  shutting  down,  the 
rest  of  the  apparatuses  water  sealed  and  the  gas  remaining  in  it  is 
kept  for  use  in  starting  up  again.  G  is  a  damper  which  is  closed 
while  the  blower  is  running.  After  the  blower  has  been  shut 
down,  the  damper  G  is  opened  and  the  air  enters  the  producer  at 
H,  passes  over  the  surface  of  the  water  in  the  vaporizer  C  and 
down  the  pipe  7,  entering  the  generator  B  at  the  bottom.  The 
pipe  J  leads  to  the  engine. 

To  operate  the  plant,  a  fire  is  lighted  just  as  in  an  ordinary  coal 
stove,  and  the  blower  is  run  until  a  good  fire  is  burning,  with  the 
relief  valve  R  open.  After  fifteen  or  twenty  minutes,  the  fire  is 
sufficiently  hot  to  give  off  gas.  The  relief  valve  is  then  closed  and 
the  gas  allowed  to  pass  through  the  apparatus  to  the  engine,  the 
blower  being  kept  running  until  the  proper  quality  of  gas  is 
obtained  at  the  test  cock  near  the  engine.  The  engine  is  then 
started,  the  blower  stopped  and  the  formation  of  gas  becomes 
automatic,  the  suction  of  the  engine  furnishing  the  draft  through 
the  fire. 

The  efficiency  of  the  gas  producer  should  be  a  little  higher  than 
that  of  a  steam  boiler.  Actual  tests  show  efficiencies  as  high  as  85 
per  cent.,  but  efficiencies  ordinarily  do  not  exceed  80  per  cent. 
The  consumption  of  fuel  in  a  gas  engine  operating  with  gas  pro- 
ducers does  not  usually  exceed  1  Ib.  per  horse-power  per  hour, 
and  in  large  installations  is  less  than  one  pound.  The  heat 
value  of  producer  gas  varies  from  100  to  150  B.T.U.  per  cubic 
foot. 


286  HEAT  ENGINES 

192.  Vaporization  of  Oil. — The  lighter  oils,  such  as  gasoline, 
are  easily  vaporized  by  either  spraying  the  oil  into  a  current  of 
air,  or  allowing  a  current  of  air  to  pass  over  the  surface  of  the  oil. 
This  vaporization  may  be  increased  or  assisted  in  four  ways: 

(a)  By  the  application  of  heat; 

(6)  By  increasing  the  surface  of  the  oil  exposed  to  the  air; 

(c)  By  reduction  of  pressure  or  increase  in  vacuum; 

(d)  By  keeping  the  air  which  is  in  contact  with  the  gasoline  as 
fresh  or  as  far  from  the  saturation  point  as  possible. 

With  the  heavier  oils,  such  as  distillates  and  crude  oil,  it  is 
necessary  to  provide  some  other  means  of  vaporizing  the  oils. 
There  are  two  general  methods  to  accomplish  this  purpose.  In 
engines  such  as  the  Hornsby-Akroyd,  the  oil  is  injected  into  a 
cylinder  against  hot  plates,  or  a  hot  ball,  and  is  almost  instantly 
vaporized  by  the  contact  with  the  red-hot  surface.  In  other 
engines  the  oil  is  vaporized  in  a  heated  chamber  external  to  the 
engine.  Initial  vaporization  is  often  produced  by  artificially 
heating  the  chamber,  and  after  the  engine  is  in  operation,  the  oil  is 
heated  by  means  of  the  exhaust  passing  through  pipes  located  in 
this  chamber.  Engines  have  been  placed  on  the  market  which 
used  crude  oil  just  as  it  comes  from  the  wells,  and  have  given  fair 
satisfaction.  The  difficulty  in  using  crude  oils  is  in  taking  care 
of  the  heavier  ingredients,  such  as  paraffine  and  asphalt,  that 
occur  in  them.  The  hot  surface  must  be  at  a  sufficient  tempera- 
ture so  that  in  vaporizing  these  heavier  oils  they  will  be  broken 
up  into  lighter  compounds  which  are  more  easily  vaporized. 
Asphalts  cannot  be  broken  up  and  must  be  removed. 

193.  Alcohol. — Alcohol  is  similar  in  its  nature  to  kerosene, 
except  that  it  will  stand  a  very  much  higher  compression,  so  that, 
while  alcohol  does  not  contain  the  heat  value  of  the  petroleum 
oils,  it  will,  nevertheless,  give  almost  as  much  power  per  pound, 
owing  to  the  fact  of  the  higher  efficiency  which  may  be  obtained 
by  its  higher  compression.     In  this  country,  alcohol  has  not  yet 
been  extensively  used,  but  it  has  been  largely  used  in  Europe  and 
Central  America.     In  using  alcohol  in  connection  with  the  engine, 
it  is  usually  necessary  to  provide  some  means  of  heating  it  so  as  to 
produce  more  rapid  vaporization.     Commercial  alcohol  usually 
contains  not  less  than  5  per  cent,  water,  and  the  percentage  may 
be  much  higher.     For  satisfactory  operation  in  a  gas  engine,  it 
should  not  contain  more  than  10  per  cent,  water. 


THE  INTERNAL  COMBUSTION  ENGINE        287 


194.  Heating  Value  of  Fuels. — The  heating  values  of  the  vari- 
ous fuels  are  given  in  the  following  table: 

TABLE  XXII. 
CALORIFIC  VALUE  OF  GASEOUS  FUELS 


Lower  heating  value 
per  cu.  ft., 
B.T.U. 

Least  air  required  for 
combustion  per  cu.  ft., 
cu.  ft. 

Oil  gas  (Pintsch)  

1000 

D  5 

Natural  gas 

950 

9  1 

Illuminating  gas  

565 

5  25 

Coke-oven  gas 

545 

5  0 

Producer  gas  (from  soft  coal)  

145 

1  25 

Producer  gas  (from  anthracite)  .... 
Producer  gas  (from  coke)  
Blast-furnace  gas  

145 
135 
100 

1.15 
1.0 

0  7 

CALORIFIC  VALUE  OF  LIQUID  FUELS 


Lower   heating 
value  per 
cu.  ft.  of  oil 
gas, 
B.T.U. 

Least  air 
required  for 
combustion 
per  cu.  ft., 
cu.  ft. 

Heating  value 
B.T.U. 
per  pound 

Heavy  crude  oil  (West  Virginia). 
Light  crude  (West  Virginia)  
Heavy  crude  (Pennsylvania)..  .  . 
Kerosene  
Gasoline 

94.6 
95.0 
99.2 
95.8" 
97  7 

IS.Olbs. 
15.0  ' 
15.0  ' 
15.0  •' 

15  0  ' 

18,320 
18,400 
19,210 
18,520 
19  000 

Benzol,  C8H6  

99  3 

13  4  ' 

17  190 

Alcohol,  100  per  cent.  .  . 

103  0 

8  6  ' 

11  664 

Alcohol  90  per  cent 

104  0 

7  8  ' 

10  080 

195.  Fuel  Mixtures. — The  mixture  of  air  and  gas  in  internal 
combustion   engines   is   very   important.     The  possible   power 
derived  from   an   engine   depends  upon   obtaining   the  proper 
mixture  of  air  and  gas.     Under  ordinary  conditions  of  pressure 
and  temperature,  a  mixture  of  CO  and  air  will  be  explosive  when 
the  range  is  from  16  to  74  per  cent.,  by  volume,  of  CO.     With 
illuminating  gas,  the  range  of  mixture  is  from  8  to  19  per  cent.; 
with  gasoline,  from  2J  to  5  per  cent.     It  will  be  noticed  that  the 
possible  range  of  mixtures  varies  very  widely  with  the  nature  of 
the   gas   used.     Experiments   show   that   the   best   results   are 
obtained  when  the  air  in  the  cylinder  is  slightly  in  excess  of  the 
theoretical  mixture. 

196.  Flame  Propagation. — A  very  important  point  in  gas  en- 
gine operation  is  the  rate  of  flame  propagation  through  the  mass 
of  the  gas.     If  this  rate  is  slow,  the  pressure  will  not  be  obtained 


288  HEAT  ENGINES 

quickly  enough  for  the  engine  to  give  its  maximum  horse-power. 
The  rate  of  flame  propagation  depends  upon  the  mixture  of  the 
gas  and  upon  the  method  of  ignition.  In  large  engines  it  is 
becoming  a  custom  to  put  more  than  one  igniter  upon  an  engine  so 
as  to  produce  more  rapid  flame  propagation.  High  compression 
has  a  tendency  to  reduce  the  rate  of  flame  propagation.  On  tfhe 
other  hand,  however,  compression  of  the  gases  increases  the  ease 
with  which  they  may  be  ignited,  and  the  range  of  the  explosive 
mixture.  Owing  to  slow  flame  propagation,  ignition  takes 
place  before  the  beginning  of  the  working  stroke. 

197.  Rated  Horse -power. — The  determination  of  the  power  of 
a  gas  engine  from  its  dimensions  is  much  more  difficult  than  of  a 
steam  engine.  The  theoretical  diagram,  although  quite  definitely 
defined  is  not  of  much  value  in  determining  the  horse-power. 
The  actual  diagram  is  influenced  by  so  many  conditions  such  as 
the  quality  and  purity  of  gas,  temperature  of  the  mixture,  condi- 
tions of  combustion,  heat  losses,  location  and  kind  of  ignition, 
form  of  combustion  chamber  and  other  items,  that  it  is  possible 
to  obtain  almost  any  result.  The  card  factor  as  applied  to  the 
steam  engine  is  of  little  value  as  it  shows  variations  under  different 
conditions  as  high  as  100  per  cent.  It  is  not  surprising  therefore 
that  numerous  methods  exist  for  determining  the.principal  dimen- 
sions of  internal  combustion  engines,  all  of  these  based  on  assump- 
tions giving  only  approximate  results. 

One  of  the  best  methods  is  based  on  the  amount  of  air  necessary 
for  combustion  and  on  the  thermal  and  volumetric  efficiencies 
when  the  engine  is  operating  with  the  quantity  of  air  assumed. 
This  method  was  developed  by  Hugo  Giildner  and  for  many 
years  this  has  been  used  with  success  for  all  sizes  and  types  of 
engines  and  for  various  fuels. 
Let         Nn  =  normal  or  rated  horse-power. 
n  =  r.p.m. 
H  =  the  lower  heating  value  of  the  fuel  in  B.T.U.— 

per  cubic  foot  for  gas — per  pound  for  liquids. 
Ch  =  fuel  consumption  per  hour  at  normal  output — 

in  cubic  feet  for  gas — in  pounds  for  liquids. 
yw  =  economic  or  thermal  efficiency  at  the  brake. 

Nn  X  33000  X  60     2545  X  Nn  ,„. 

rhen'      *•  =   ^  ~ 


„        2545  X  N, 
whence  G&  =       — ^rw~ 
rjw  A /i 


THE  INTERNAL  COMBUSTION  ENGINE        289 

Let  Cst  =  fuel  consumption  per  suction  stroke. 
Lst  =  air  comsumption  per  suction  stroke. 
L  =  proper  amount  of  air  in  cubic  feet  required  per  cubic 

foot  of  gas  or  per  pound  of  liquid  fuel. 
D  =  diameter  of  the  cylinder. 
S  =  stroke  of  the  engine. 
t]v  =  volumetric  efficiency. 
Then,  for  a  single  acting  four-cycle  engine 

Ch         ^.SXN^ 


ChXL      84.8XJVnXL 
=  30  Xn~    nXHXrjw 

For  two-cycle  engines  equations  (27)  and  (28)  must  be  divided  by  2  . 
During  one  suction  stroke  the  volume  of  the  actual  charge 
drawn  in  is 

_  r         T     _  D2  X  ?r  X  S  _  actual  piston  displacement 
C.«  +  Lst  -        4  x  ^-  volumetric  efficiency 

84.8XATnX(l+L)   .        .... 
n  -        m  Cublc  feet 


Solving  for  D,  S,  and  n,  we  get  for  engines  using  gaseous  fuels: 


=  Vsx     SfS^-A  (») 


„  n  f 

X^X^ 
(1+L) 

Xr,v  ^'m'  (32) 

For  liquid  fuel  engines  the  term  (1-f-  L)  may  be  put  equal  toL, 
as  the  volume  of  fuel  is  very  small  compared  with  the  volume  of 
air. 

In  view  of  the  amount  of  experimental  data  available  a  selec- 
tion of  the  efficiencies  rjw  and  t]v,  and  the  air  consumption  for 
different  types  and  sizes  of  engines  can  be  easily  made.  For  this 
purpose  Tables  XXIII  and  XXIV  are  inserted. 

TABLE  XXIII.— VOLUMETRIC  EFFICIENCY— rjv,  OF  GAS  ENGINE 
rjv  =  .88  —    .93    For  slow  speed  engines  with  mechanically  operated  inlet 

valve. 
?7v  =  .80  —    .87    For  slow  speed  engines  with  automatically  operated  inlet 

valves. 
rjv  =  .78  —    -85    For  high  speed-  engines  with  mechanically  operated  inlet 

valve. 

?7v  =  .65  —    .75    For  high  speed  engines  with  automatic  inlet  valve. 
r)v  =  .50  —    .65    For  very  high  speed  automobile  engines  with  automatic 

inlet  valves  and  air  cooling. 
19 


290 


HEAT  ENGINES 


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THE  INTERNAL  COMBUSTION  ENGINE        291 

Diameter  D,  stroke  S  and  speed  n  have  a  certain  relation  to  each 

o 

other.     Present-day  engines  can  safely  be  built  with  ^  =  1  to  2.5, 

and  with  piston  speeds  up  to  800  ft.  per  minute  for  large  engines 
and  1200  for  smaller  ones. 

Example. — Determine  the  cylinder  diameter,  stroke  and  revolutions 
per  minute  of  a  four-cycle,  single-acting,  one-cylinder  anthracite  pro- 
ducer gas  engine  of  170  H.P.,  having  a  piston  speed  of  800  ft.  per 

minute  and  a  stroke-diameter  ratio  -w  =  1.35 

Solution.— From  table  XXIII  use  rjv  =  .90     T 
From  table  XXIV  use  17.  =.26 

H  =  140  B.T.U.  and  since  n  =  ^  =  ^= 

Zo         Z.IL) 

solving  for  D 


I     W8XNn(l+L)       =      I 
\S  XnXHXirvX  77,       V1-3 


108  X  170  X  2.5  X  2.7 


.35  X  800  X  140  X  .26  X  .9 
=  1.877'  feet  =  22.5  inches 
S  =  1.35  X  22.5  =  30  inches 

800 

77  =  2~X30  =        ;  r'P'm> 

Example. — Determine  the  diameter  and  stroke  of  a  two-cycle, 
single-acting,  4-cylinder  Diesel  Oil  Engine  of  150  H.P.,  having  350 
r.p.m.  and  a  piston  speed  of  700  ft.  per  minute. 

Solution. — 

From  table  XXIII  rjv  =  -80 

From  table  XXIV  i,w  =  .32  L  ~  18'°° 

Each  cylinder  must  develop  2       .  =  18.75  H.P.  per  cycle. 

700  X  12 
*  -  2  X  350  = 


108  X  18.75  X  312  _  7    „ 

--  =  /  .o 


X  350  X  18,000  X  .32  X  0.80 

The  American  Association  of  Automobile  Manufacturers 
determines  the  normal  output  of  four-cycle  automobile  engines  by 
the  formula — 

B.H.P.  =    ^|y^  (33) 

where  d  =  the  diameter  of  the  cylinder  in  inches,  and  N  the  num- 
ber of  cylinders.     This  rule  is  based  on  a  piston  speed  of  1000  ft. 


292  HEAT  ENGINES 

per  minute  and  has,  of  course,  an  arbitrary  and  conventional  value 
only. 

The  rated  horse-power  of  a  gas  engine  to  drive  a  given  size 
electric  generator  is  quite  different  from  that  of  a  steam  engine 
to  drive  the  same  machine.  This  is  due  to  the  fact  that  a  gas 
engine  as  rated  has  very  little  overload  capacity,  while  a  steam 
engine  can  carry  a  25  per  cent,  overload  continuously  and  a 
50  per  cent,  overload  for  a  short  period  of  time.  In  order  to 
allow  for  the  overload  capacity  of  the  generator,  the  gas  engine 
must  be  sufficiently  large  to  drive  the  generator  under  that 
condition. 

As  an  example,  to  drive  a  2000  k.w.  generator,  a  4500  H.P. 
gas  engine  is  used,  while  to  drive  the  same  generator  with  a 
steam  engine,  a  3000  H.P.  engine  is  used. 

It  should  be  noted  that,  at  present,  gas  engines  are  rated 
on  their  output,  or  brake  horse-power,  while  steam  engines  are 
rated  on  their  indicated  horse-power,  and  that,  as  stated  above, 
gas  engines  are  rated  at  practically  their  maximum  capacity, 
while  steam  engines  are  rated  at  the  I. H.P.  at  which  they  give 
the  best  economy. 

198.  Actual  Horse -power. — The  actual  indicated  horse- 
power (I. H.P.)  of  a  gas  engine  already  built  and  in  operation 
may  be  determined  in  almost  exactly  the  same  way  as  was 
done  in  the  case  of  the  steam  engine,  the  only  difference  being 
that  in  the  formula, 

plan 
i.tt.r.  -  33000' 

n  =  explosions  per  minute,  when  finding  the  horse-power  of 
the  gas  engine,  while  when  finding  the  power  of  the  steam 
engine,  it  was  equal  to  the  revolutions  per  minute. 

In  both  cases,  I  =  the  length  of  stroke  in  feet,  and 

a  =  the  cross-sectional  area  of  the  piston  in 
square  inches. 

The  mean  effective  pressure  p  is  found  by  taking  indicator 
cards  from  the  engine  and  then  multiplying,  by  the  scale  of 
the  spring  used,  the  quotient  found  by  dividing  the  area  of 
the  card  by  its  length. 

The  cross-sectional  area  of  the  piston  in  the  gas  engine  indi- 
cator is  usually  one-fourth  of  a  square  inch,  while  that  of  the 


THE  INTERNAL  COMBUSTION  ENGINE        293 

steam-engine  indicator  is  one-half  a  square  inch,  the  difference 
being  due  to  the  fact  that  the  initial  pressure  in  the  gas-engine 
cylinder  is  so  much  greater. 

The  brake  horse-power  (B.H.P.)  of  a-  gas  engine  is  found  in 
exactly  the  same  manner  as  the  B.H.P.  of  a  steam  engine,  the 
expression  being 

(35> 


where  I  =  the  length  of  the  brake  arm  in  feet, 
w  =  the  net  weight  on  the  brake,  and 
n  =  the  number  of  revolutions  per  minute. 

It  is  thus  seen  that  in  making  a  test  of  a  gas  engine  to  obtain 
the  I.H.P.  and  B.H.P.,  both  the  explosions  per  minute  and 
the  revolutions  per  minute  must  be  noted. 

Example.  —  A  10|"  X  16|"  single-acting  gas  engine  runs  200  r.p.m. 
and  makes  96  explosions  per  minute.  The  gross  weight  on  the  brake 
was  140  Ibs.,  the  tare  20  Ibs.,  and  the  length  of  the  brake  arm,  60  in. 
The  area  of  the  indicator  card  was  1.07  sq.  in.  and  the  length  3  in., 
and  the  scale  of  the  spring  used  was  219  Ibs.  Find  the  (a)  I.H.P.; 
(6)  B.H.P.;  (c)  F.H.P.;  and  (d)  mechanical  efficiency. 

Solution.  — 

(a)    M.E.P.  =  ~'^  X  219  =  78.1  Ibs. 
o 

a  =  TT  X  5T3F  X  5f\  =  84.5  sq.  in. 

I  =  16J  -5-  12  =  1.406  ft. 

plan   _  78.1  X  1.406  X  84.5  X  96  _  885000 
I.M.I  .  -  330()()  -  =  -  27. 


(6)  Net  weight  =  140  -  20  =  120  Ibs. 

Length  of  brake  arm  =  60  4-  12  =  5  ft. 

2irlnw       2  X  3.1416  X  5  X  200  X  120 


B.H.P.  = 


33000  33000 

_  755000 
"33000    • 


(c)  F.H.P.  =  I.H.P.  -  B.H.P.  =  27  -  22.85  =  4.15. 

(d)  Mech.  Eff.  =      '-       =         ^"  =  -846  =  84-6  Per  cent- 


CHAPTER  XVII 


DETAILS  OF  GAS-ENGINE  CONSTRUCTION 

199.  In  general,  the  frame  and  working  parts  of  the  gas  engine 
are  heavier  in  construction  than  the  corresponding  parts  of  a 
steam  engine.  This  is  largely  due  to  the  fact  that  the  number  of 
impulses  given  the  gas  engine  for  the  same  power  is  less  than 
those  given  the  steam  engine,  and  hence  each  impulse  in  the  gas 
engine  must  exert  more  force. 

FRAME. — Fig.  174  shows  the  frame  of  a  modern  gas  engine  of 
medium  size.  The  barrel  of  the  cylinder  is  cast  with  the  frame. 
The  main  bearing  supports  are  cast  in  the  same  frame. 


FIG.  174. — Gas  engine  frame. 

CYLINDER  AND  PISTON. — The  inner  lining  of  the  cylinder  is 
inserted  in  the  frame  as  a  separate  piece,  except  in  the  smaller 
engines. 

Fig.  175  shows  the  piston  and  piston  rings.  Three  rings,  at 
least,  and  often  six  or  seven,  are  used  in  a  gas  engine.  It  is  very 
important  that  the  piston  fit  the  cylinder  as  closely  as  possible  so 
as  to  hold  the  compression.  The  piston  shown  is  for  a  single- 

294 


DETAILS  OF  GAS-ENGINE  CONSTRUCTION     295 


acting  engine,  and  serves  both  as  piston  and  cross-head.  The 
cross-head  pin  is  shown  at  the  top  of  the  figure,  and  is  placed  in 
the  hole  shown  in  the  side  of  the  piston.  This  is  the  most  com- 
monly used  construction  for  small  and  medium  size  engines. 

CONNECTING  RODS. — The  connecting  rods  used  in  gas  engines 
are  similar  to  those  in  steam-engine  practice. 

VALVE  MECHANISM. — The  valves  used  have  been  almost  the 
same  for  all  types  of  gas  engines,  and  are  of  the  poppet  type. 
The  exhaust  valves  are  always  mechanically  operated,  but  the 
inlet  valves  may  be  either  automatic  or  mechanically  operated. 

Fig.  176  shows  the  cross-section  of  a  four-cycle  gas  engine, 
and  shows  both  inlet  and  exhaust  valves.  These  valves  are 
operated  from  a  cam  shaft  at  the  side  of  the  engine  by  means 
of  roller  cams.  In  some  engines  these  cams  are  replaced  by 
eccentrics. 


FIG.  175. — Piston  and  rings  for  gas  engine. 

WATER-JACKET. — -In  all  except  small  air-cooled  engines,  the 
cylinder  and  cylinder  head  are  cooled  by  being  surrounded  by 
a  water-jacket,  and  in  the  best  engines  the  valves  are  also  water- 
jacketed.  The  water-jackets  are  shown  in  Fig.  176,  surrounding 
the  valves  and  reaching  between  the  valves. 

200.  Ignition. — One  of  the  most  important  details  of  gas- 
engine  construction  has  been  the  development  of  a  suitable 
means  of  ignition.  The  first  successful  form  of  ignition  was  by 
means  of  an  open  flame  which  was  drawn  into  the  cylinder  at 
the  proper  time.  Flame  ignition,  however,  is  uncertain  and 
difficult  of  application,  and  is  not  economical  and  so  has  been 
abandoned  in  recent  engines. 

The  next  form  of  ignition  was  the  hot  tube,  in  which  a  closed 
tube  connected  with  the  engine  cylinder  was  kept  at  red  heat 


296 


HEAT  ENGINES 


by  means  of  an  external  flame.  The  compression  of  the  gases 
into  the  hot  tube  ignites  them  at  the  proper  time  in  the  stroke. 
The  time  of  ignition  is  more  or  less  regulated  by  the  tempera- 
ture of  the  tube.  In  some  cases  the  admission  of  the  gas  into 
the  hot  tube  was  controlled  by  a  valve.  This  form  of  ignition 


is  satisfactory  in  small  engines,  but  is  hardly  sufficient  to  ignite 
a  large  volume  of  gas  such  as  is  admitted  to  a  large  engine,  and 
does  not  admit  of  a  change  in  the  time  of  sparking. 

One  of  the  simplest  forms  of  igniters  is  that  used  by  the  Deisel 
Engine  Company.     In  this  engine  the  air  is  compressed  to  a  very 


DETAILS  OF  GAS-ENGINE  CONSTRUCTION     297 

high  pressure  and  the  temperature  is  then  sufficient  to  ignite 
the  entering  charge  of  oil,  or  gas,  which  is  delivered  to  the  cylinder 
at  a  pressure  slightly  higher  than  the  compression  pressure. 
This  then  requires  no  special  igniting  apparatus,  and  the  time 
of  ignition  is  controlled  by  the  time  of  admission  to  the  cylinder. 
In  the  Hornsby-Akroyd  oil  engine,  a  hot  bulb,  used  for  vaporizing 
the  entering  oil,  serves  also  as  an  igniter. 

At  the  present  time,  the  most  used  and  the  most  successful 
form  of  ignition  is  by  electric  spark.     This  has  proven  to  be  the 


FIG.  177. — Magneto. 

most  satisfactory  in  the  large  majority  of  internal  combustion 
engines  and  in  automobiles  it  is  used  exclusively.  It  is  by  far 
the  most  reliable  and  flexible  method  in  use.  There  are  various 
means  of  generating  the  current  used  in  electrical  ignition,  such 
as  a  battery,  dynamo,  or  magneto,  the  one  most  commonly 
employed  at  present  being  the  magneto,  Fig.  177.  This  is  a  piece 
of  apparatus  consisting  of  a  permanent  steel  magnet  bent  in  the 
form  of  a  letter  U,  with  a  coil  of  wire  revolving  in  the  opening 
between  the  poles  or  ends  of  the  magnet.  As  the  wire  revolves 
it  cuts  the  magnetic  lines  of  force  and  an  electric  current  is  set 
flowing  through  the  wire. 

When  the  armature  is  in  a  position  such  that  the  magnetic 
lines  flow  through  the  core,  this  core  becomes  magnetized.  If 
the  armature  is  then  turned  so  that  the  lines  no  longer  flow  through 


298 


HEAT  ENGINES 


the  core,  the  core  loses  its  magnetism  and  it  is  this  dying  away 
of  the  magnetism  of  the  core  that  produces  an  electric  current 
in  the  winding. 

A  magneto  thus  gives  a  current  only  at  certain  points  in  the 


FIG.  178.— Bottom  of  make- 
and-break  jignitor  block,  show- 
ing contact  points. 


FIG.  179. — Top  of  make-and-break 
ignitor  block. 


revolution  of  the  armature  and  it  must  be  so  driven  that  it  will 
give  this  current  at  the  time  the  engine  needs  a  spark. 

The  more  suddenly  the  magnetism  changes  strength  the  more 
intense  the  current  will  be.  If  the  armature  is  turned  slowly, 
the  current  may  not  be  strong  enough  to  form  a  spark. 


FIG.  180. — Section    of    cylinder    head    showing    make-and-break    ignition 

system. 

A  battery  of  dry  cells  is  frequently  used  to  furnish  current  for 
ignition  of  small  engines. 

There  are  two  forms  of  electric  ignition;  viz.  the  make-and- 
break  or  low  tension,  and  the  jump  spark  or  high  tension  systems. 
Of  these,  the  make-and-break  is  the  simpler. 

In  the  make-and-break  system  there  are  two  contact  points, 
as  shown  in  Figs.  178  and  180,  located  inside  of  the  cylinder,  and 


DETAILS  OF  GAS-ENGINE  CONSTRUCTION     299 

in  addition,  in  series  with  the  circuit  is  placed  what  is  called  a 
spark  coil.  This  consists  of  a  number  of  turns  of  comparatively 
heavy  wire  wrapped  around  a  core  composed  of  iron  wires. 
This  coil  acts  as  an  inductive  resistance,  and  when  the  circuit 
is  broken  it  serves  to  cause  a  hot  spark  at  the  point  of  the  break. 
The  circuit  of  the  make-and-break  igniter,  then,  consists  of  a 
battery,  or  magneto,  and  a  spark  coil,  both  of  which  are 
placed  in  series  with  two  contact  points  in  the  engine.  Just 
before  the  point  of  sparking,  the  two  contact  points  A  and  B, 
Fig.  178,  are  brought  together,  and  at  the  point  of  sparking 
the  mechanism  is  so  constructed  (see  Fig.  179)  that  the  two 


FIG.  181. — Diagram  of  jump-spark  ignition. 

points  are  quickly  separated,  producing  a  sufficient  spark  to 
ignite  the  charge. 

The  advantages  of  the  make-and-break  system  are:  (a)  hot 
spark  ensuring  ignition;  (6)  little  trouble  with  insulation.  The 
disadvantages  are :  (a)  moving  mechanism  required  in  the  cylinder; 
(6)  points  of  contact  become  foul  and  wear  away. 

The  make-and-break  igniter  is  used  in  a  great  many  engines, 
and  is  advocated  by  many,  owing  to  the  low  tension  at  which  it 
is  operated.  It  is  the  most  common  form  of  ignition  on  station- 
ary engines. 

In  jump-spark  ignition,  Figs.  181  and  182,  the  current  is  taken 
from  a  battery  B,  Fig.  181,  or  generator  at  a  low  voltage  and  passed 
through  an  induction  coil  C,  having  an  interrupter.  The  induction 
coil  has  a  primary  and  secondary  coil.  The  interrupted  current 
passing  through  the  primary  coil  induces  a  high-tension  current 
in  the  secondary  coil.  This  current  at  a  high  voltage  is  carried  to 
what  is  known  as  a  spark  plug  E,  located  in  the  engine  cylinder. 


300 


HEAT  ENGINES 


This  spark  plug  contains  two  points  about  ¥V  m-  apart,  across 
which  a  high-tension  current  is  made  to  j  ump  at  the  time  of  igni- 
tion. The  time  of  ignition  is  controlled  by  a  timer  D,  fastened  to 
the  engine  shaft,  and,  at  the  proper  time  of  the  stroke,  this  timer 
closes  the  battery  circuit,  the  high-tension  current  is  generated  in 
the  induction  coil,  and  the  spark  jumps  across  the  air  gap  causing 
ignition  in  the  cylinder.  There  are  a  great  many  detailed  modifi- 
cations of  this  device,  but  the  above  description  covers  the  general 
construction  of  them  all.  In  some  cases  the  current  is  furnished 
by  an  alternating-current  magneto.  With  an  alternating  current, 
no  interrupter  is  necessary.  This  system  is  almost  universally 
used  on  automobiles. 

The  advantages  of  the  jump-spark  system  are:  (a)  absence  of 
moving  parts  in  the  cylinder;  (b)  easy  adjustment  of  the  time  of 


FIG.  182. — Section  of  cylinder  head  showing  jump-spark  ignition  system. 

ignition.  The  disadvantages  are:  (a)  high  insulation  required; 
(b)  liability  of  spark  plug  becoming  fouled  with  oil  or  dirt;  (c) 
intensity  of  spark  varies  with  pressure  in  cylinder. 

In  all  forms  of  gas-engine  igniters,  some  means  should  be  pro- 
vided for  changing  the  time  of  ignition,  so  that  the  pressure  may 
reach  a  maximum  at  the  proper  time  in  the  stroke.  In  the  jump- 
spark  igniter  this  is  done  by  moving  the  position  of  the  commuta- 
tor relative  to  the  piston  position.  The  proper  time  for  ignition 
depends  upon  the  mixture  and  the  speed  of  the  engine. 

Ignition  is  not  instantaneous  and  in  order  to  have  the  greatest 
pressure  against  the  piston  when  it  begins  the  power  stroke,  the 
mixture  must  be  set  on  fire  before  the  completion  of  the  compres- 
sion stroke.  This  is  called  advancing  the  spark. 

201.  Governing. — The  aim  of  all  governors  is  to  obtain  the 
maximum  thermal  efficiency  at  all  loads.  The  governing  of  a 
gas  engine  is  different  from  that  of  a  steam  engine.  In  a  steam 
engine  under  a  constant  load,  each  cycle  of  the  engine  is 


DETAILS  OF  GAS-ENGINE  CONSTRUCTION     301 

practically  the  same,  while  in  the  gas  engine,  even  with  a 
constant  load,  there  is  always  some  change  in  the  cycle  of  the 
engine.  This  is  due  to  changes  of  mixture  and  time  of  ignition. 
This  makes  the  problem  of  governing  in  the  gas  engine  more  diffi- 
cult than  in  the  steam  engine. 

The  following  general  methods  of  governing  are  used  in  gas 
engines: 

I.  The  "hit  and  miss"  system. 

II.  Variation  in  the  quantity  of  charge  entering  the  cylinder,  the 
mixture  of  gas  and  air  being  constant. 

III.  Variation  of  the  mixture  of  gas  and  air,  the  load  determin- 
ing the  quality  of  the  mixture. 

IV.  Governing  by  changing  the  time  of  ignition. 

V.  Combinations  of  the  above  methods. 

1.  Hit  and  Miss. — The  most  common  of  all  these  systems  of 
governing  is  the  "hit  and  miss."     In  this  form  of  governing, 
when  the  speed  exceeds  the  normal,  the  supply  of  gas  is  cut  off 
and  the  engine  gets  no  explosion,  causing  the  engine  to  "miss." 
The  loss  of  the  explosion  causes  the  speed  to  slacken,  the  governor 
opens  the  inlet  valve  and  the  engine  again  receives  an  impulse,  or 
a  "hit." 

This  is  most  economical  and  simplest  method  of  governing,  but 
does  not  provide  the  closest  regulation  in  speed.  In  this  system 
the  "miss"  may  be  occasioned  by  (a)  holding  the  exhaust  valve 
open  and  thus  allowing  no  suction,  or  (b)  by  failing  to  open  the 
gas  valve.  There  may  be  considerable  variation  in  speed. 

This  method  of  governing  is  not  desirable  for  large  engines 
because  of  the  high  pressures  after  a  "miss." 

2.  Quantity  governing  may  be  accomplished  by  varying  the 
weight,  or  quantity,  of  the  mixture  of  gas  and  air  entering  the 
cylinder.     This  result  may  be  obtained  in  two  ways. 

1.  By  cutting  off  the  charge  before  the  piston  reaches  the  end 
of  the  suction  stroke. 

2.  By  throttling  the  charge  during  the  suction  stroke. 

The  disadvantage  of  this  system  is  that  the  compression  varies 
with  the  size  of  the  charge.  Reducing  the  compression  reduces 
the  efficiency,  and  hence  this  form  of  governing  is  not  as  univer- 
sally efficient  as  the  "hit  and  miss." 

3.  Quality   Governing  .—In    this     system   the  weight  of  the 
charge  remains  the  same,  but  the  proportion  of  gas  to  air  is 
varied — the  governor    usually    controlling  the    supply    of   gas. 


302  HEAT  ENGINES 

As  the  load  decreases,  the  amount  of  gas  is  reduced  for  the  same 
total  charge.  This  system  has  the  advantage  over  Method  No. 
2,  that  the  pressure  of  the  compression  always  remains  the  same. 
On  light  loads,  however,  it  is  not  so  economical  as  Method  No.  2, 
for  when  the  load  is  very  light  the  mixture  may  be  so  weak 
that  the  charge  will  not  ignite. 

Method  No.  4. — Controlling  the  speed  by  changing  the  time  of 
ignition  is  used  on  automobile  engines.  As  the  load  diminishes, 
the  time  of  sparking  is  brought  nearer  to  the  working  stroke, 
that  is,  it  is  advanced,  and  it  may  even  occur  after  the  dead 
center  (just  previous  to  the  working  stroke).  As  the  spark  is 
advanced,  the  engine  develops  less  and  less  power.  The  quan- 
tity and  quality  of  the  charge,  however,  remains  the  same.  This 
system  of  speed  control  is  very  uneconomical  at  light  loads. 

Method  No.  5. — A  great  many  different  combinations  of  the 
above  systems  have  been  used.  Often  engines  having  "quan- 
tity" and  "quality"  governors  for  the  heavy  and  medium  loads 
change  the  governing  system  to  "  hit  and  miss  "  for  light  loads.  A 
combination  largely  used  in  electric  lighting  work,  on  account  of 
the  close  regulation  obtained,  is  quality  governing  at  high  loads 
and  quantity  governing  at  low  loads. 

The  governing  of  an  automobile  is  a  combination  of  quality 
governor  by  the  throttle,  and  governing  by  spark  advance  with 
the  ignition  device. 

Kerosene  and  fuel  oil  engines  are  commonly  governed  by  by- 
passing the  fuel  so  that  a  greater  or  less  amount  of  it  is  injected 
into  the  cylinder. 

Gas-engine  governing  is  at  present  almost  as  perfect  as  govern- 
ing in  the  steam  engine.  There  is  no  difficulty  in  obtaining  suffi- 
ciently accurate  governing  so  that  alternators  driven  by  gas 
engines  may  be  operated  in  parallel. 

202.  Carburetors. — A  carburetor  is  a  device  used  for  vaporizing 
oil,  particularly  gasoline.  It  is  largely  used  in  connection  with 
automobile  or  small  launch  engines.  In  a  carburetor  the  air 
may  be  passed  over  or  through  the  gasoline,  or  the  gasoline  may 
be  mechanically  sprayed  into  the  current  of  incoming  air. 

Fig.  183  shows  a  cross-section  of  one  type  of  Stromberg  carbu- 
retor. A  float  M  operates  a  pair  of  levers  and  through  them  the 
needle  valve  K,  thus  controlling  the  supply  of  gasoline  to  the 
spray  nozzle  C.  The  gasoline  enters  the  carburetor  from  the 
source  of  supply  through  0  and  the  dirt  in  it  is  removed  by  the 


DETAILS  OF  GAS-ENGINE  CONSTRUCTION     303 


strainer  N.  The  hot  water  in  the  jacket  J  keeps  the  carburetor 
warm  and  assists  in  vaporizing  the  gasoline. 

At  each  charging  stroke  of  the  engine,  air  is  drawn  in  through 
the  fixed  air  inlet  S  and  passes  at  a  high  velocity  up  through  the 
venturi  tube  D  and  around  the  nozzle  C.  The  gasoline  is  sucked 
in  a  jet  from  this  nozzle  and  is  mixed  with  the  air  in  the  mixing 
chamber  1.  The  throttle  valve  H  regulates  the  supply  of  the 
mixture  to  the  engine. 

As  the  speed  of  the  engine  increases,  the  proportion  of  air  to 
gasoline  must  be  increased.  This  is  taken  care  of  by  the  auxiliary 


\ 


EXPLANATION 


A- Low  .speed  adjusting  nut     K- Needle  valve 

B-High  speed  adjusting  nut    L- Glass  float  chamber 

C  -  Spray  nozzle 

D- Venturi  tube 

E-  Auxiliary  air  valve 

F-  Low  speed  spring 

G-High  speed  spring 

H- Throttle  valve 

I  -  Mixing  chamber 

J- Water  jacket 


M-  Metal  float 

N-  Gasoline  strainer 

O  -  Gasoline  line  coupling 

P  -  Drain  cock 

Q-  Hot  air  horn 

R-  A.ir  shut-off  for  starting 

S  -  Fixed  air  inlet 

T-  Season  adjustment 


FIG.  183. — Cross-section  of  Stromberg  gasoline  carburetor. 

air  valve  E,  which  is  opened  or  closed  a  greater  or  less  amount 
as  the  speed  of  the  engine  increases  or  decreases. 

203.  Vertical  Versus  Horizontal  Engines. — The  advantages  of 
the  vertical  engine  are:  higher  rotative  speed,  better  balancing, 
occupy  less  floor  space,  less  wear  on  the  cylinders  and  pistons. 
The  advantages  are  obtained  because  multi-cylinder  engines  are 
more  easily  built  of  the  vertical  type.  Therefore  more  cylinders, 
each  of  smaller  size,  may  be  used  for  the  same  power  than  would 


304 


HEAT  ENGINES 


be  the  case  with  a  horizontal  engine.  This  means  a  shorter  stroke 
and  hence  higher  rotative  speed  for  the  same  power.  An  increase 
in  the  number  of  cylinders  means  better  balancing  and  less  vibra- 
tion. The  reciprocating  masses  of  the  horizontal  engine  tend 
to  cause  the  engine  to  move  on  its  foundation  and  heavier  founda- 
tions are  necessary  than  in  case  of  vertical  engines. 

The  disadvantages  of  the  vertical  engine  are:  increased  first 
cost,  and,  in  the  enclosed  type,  too  much  oil  may  get  in  the  cylin- 
der, causing  trouble.  The  open  end  of  the  cylinder  on  the  hori- 
zontal engine  assists  in  cooling  the  piston.  In  the  larger  size 


FIG.  184. — Koerting  two-cycle  gas  engine. 


engines,  the  cylinders  are  generally  horizontal,  while  most  auto- 
mobile and  small  launch  engines  have  vertical  cylinders. 

204.  Large  Gas  Engines. — In  the  large  sizes,  the  single-acting 
engine  has  been  replaced  by  the  double-acting  engine,  similar 
in  its  arrangement  to  the  steam  engine.  Fig.  184  shows  a  block 
plan  of  a  modern  two-cycle  gas  engine  of  the  double-acting  type. 
In  this  figure,  the  device  for  cooling  the  piston  and  piston  rod  is 
not  shown.  In  most  large  engines,  however,  of  the  double- 
acting  type,  the  piston  and  piston  rod  are  cooled  by  allowing 
a  circulation  of  water  through  them.  Usually  the  water  enters 
through  a  flexible  pipe  connected  to  the  cross-head,  and  is  re- 
moved by  a  tail  rod  projecting  through  the  cylinder  head. 


DETAILS  OF  GAS-ENGINE  CONSTRUCTION     305 

205.  Oil  Engines  for  Ships. — For  use  in  marine  work  certain 
conditions  are  required  for  successful  operation  of  the  engine. 

1.  It  should  be  able  to  be  started  quickly  from  any  position 
without  having  to  be  barred  round. 

2.  It  should  be  capable  of  rapid  reversal. 

3.  It  should  be  able  to  run  continuously  for  long  periods  with- 
out a  stop. 

4.  It  should  work  economically  at  various  speeds. 

5.  It  should  start  under  a  load. 

6.  It  should  admit  of  easy  inspection  and  adjustment. 

7.  It  should  work  smoothly  in  a  rough  sea  when  the  propeller 
is  sometimes  partly  out  of  the  water. 

The  relative  advantages  and  disadvantages  of  Diesel  engines 
as  compared  with  steam  engines  and  boilers  for  use  on  large  ships 
are  as  follows: 
Advantages : 

(a)  Have  much  higher  thermal  efficiency; 

(6)  Weigh  about  half  as  much  and  occupy  about  two-thirds 
the  space  for  the  same  power; 

(c)  Make  possible  cleaner,  quicker  and  easier  "coaling;" 

(d)  Require  less  attendants; 

(e)  Eliminate  funnels  and  dirt; 

(f)  Start  quicker; 

(g)  Eliminate  stand-by  losses. 
Disadvantages. 

(a)  Are  not  so  easily  reversed  or  maneuvered  in  harbors; 

(b)  Fuel  is  more  expensive  in  most  places  and  not  so  readily 
available ; 

(c)  There  is  an  absence  of  steam  for  working  the  auxiliary 
devices. 

206.  Humphrey  Gas  Pump— During  the  Brussels  Exposition 
in  1910  there  was  exhibited  a  new  type  of  pumping  engine  known 
as  the  Humphrey  Gas  Pump.     Since  that  time  this  gas  pump  has 
gained  a  world-wide  reputation.     It  has  been  successfully  intro- 
duced into  this  country  and  has  been  greatly  improved  by  Ameri- 
can designers.     The  largest  pump  in  the  world,  pumping  water 
for  the  City  of  London,  is  of  this  type.     Its  operation  can  best 
be  understood  by  reference  to  Fig.  185. 

The  pump   consists  of  a  vertical  gas  cylinder  A  with  inlet 
and  outlet  valves  B  and  C.     These  valves  interlock  with  each 
other.     On  the  water  side  of  the  pump  there  is  a  suction  pipe  D, 
20 


306 


HEAT  ENGINES 


a  suction  valve  S,  and  a  pressure  pipe  E  connecting  the  cylinder 
with  the  pressure  tank  F.  The  water  column  G  forms  a  gas-tight 
piston.  The  operation  of  the  pump  is  as  follows: 

We  will  assume  at  the  beginning  that  the  gas  cylinder  is  filled 
with  a  mixture  of  gas  and  air.  This  charge  of  gas  and  air  is 
ignited  and  the  pressure  is  suddenly  increased.  While  this 
takes  place  the  volume  will  scarcely  change  so  that  combustion 
practically  takes  place  at-  constant  volume.  The  water  column 


FIG.  185. — Diagram  of  Humphrey  gas  pump. 

owing  to  the  increased  pressure  on  its  surface  is  rapidly  acceler- 
ated by  the  pressure  in  the  gas  cylinder  and  the  gases  undergo 
adiabatic  expansion.  When  the  gas  has  reached  a  predetermined 
pressure  the  exhaust  valves  on  the  top  of  the  cylinder  and  the  suc- 
tion valves  on  the  water  inlet  begin  to  open  automatically.  The 
inflowing  water  follows  the  moving  water  column  and  fills  the  gae 
cylinder,  replacing  the  burned  gases.  The  hydrostatic  pressurs 
from  the  water  tank  reverses  the  water  column  closing  the  water 
inlet  valves  and  forcing  out  the  bur-ned  gases  through  the  exhaust 
valve.  When  the  water  level  reaches  the  position  V3,  the  exhaust 
valve  closes  and  the  water  column  compresses  the  remaining 
burned  gases  to  the  volume  V*.  Now  the  water  column  reverses 


DETAILS  OF  GAS-ENGINE  CONSTRUCTION      307 

again.  Reexpansion  of  the  compressed  gases  takes  place  and 
the  pressure  falls  below  the  atmospheric  pressure.  The  mixing 
inlet  valve  opens  and  the  new  charge  is  taken  in  until  the  volume 
Vi  is  filled.  The  water  column  again  reverses  and  compresses 
the  charge  to  the  volume  Vz-  Ignition  lakes  place  and  the  whole 
cycle  is  started  over  again.  The  engine  works  on  the  four-cycle 
principle,  the  expansion  and  contraction  occurs  adiabatically  and 
the  cycle  is  carried  on  by  the  oscillation  of  the  water  column  due 
to  the  changes  of  pressure.  The  action  of  the  pump  is  not  altered 
if  instead  of  delivering  into  the  elevated  tank  it  is  discharged  into 
an  open  air  vessel  or  into  an  open  tower.  The  pump  has  the 
advantage  of  being  capable  of  handling  enormous  quantities  of 
water.  In  the  large  pump  installed  in  the  City  of  London,  15 
tons  of  water  are  discharged  at  each  discharge  of  the  pump,  the 
pump  having  a  capacity  of  150  million  gallons. 

PROBLEMS 

1.  A  gasoline  engine  uses  1  Ib.  of  gasoline  per  I.H.P.  per  hour.     If  the 
gasoline  contains  19,500  B.T.U.  per  pound,  what  is  the  actual  heat  emciency 
of  the  engine? 

2.  A  gas  engine  uses  20  cu.  ft.  of  gas  per  horse-power  per  hour.     Each 
cubic  foot  of  gas  contains  600  B.T.U.     Initial  temperature  in  the  engine 
is  2000°  and  the  final  temperature  800°.     What  is  the  actual  and  theoretical 
thermal  efficiency  of  the  engine? 

3.  What  is  the  mechanical  efficiency  of  an  8\"  X  14"  single-acting  gas 
engine  if  it  runs  225  r.p.m.,  makes  106  explosions  per  minute,  has  a  net  weight 
of  50  Ibs.  on  the  brake,  and  the  M.E.P.  is  76.8  Ibs.  ?     The  length  of  the  brake 
arm  is  62.75  in.  and  the  tare  of  the  brake  is  19  Ibs. 

4.  A  card  from  an  8j"X  14",  single-acting  gas  engine  has  an   area  of 
.9  sq.  in.  and  its  length  is  3  in.     Scale  of  spring,  240  Ibs.;  r.p.m.,  225. 
Explosions  per  minute,  100.     There  is  a  Prony  brake  on  the  engine,  the 
length  of  the  brake  arm  being  63  in.  and  the  net  weight  on  the  brake  42  Ibs. 
Find  the  I.H.P. ;  B.H.P.;  F.H.P.;  and  the  mechanical  efficiency. 

6.  An  8"X  10",  single-acting  steam  engine  running  250  r.p.m.  and 
having  an  average  M.E.P.  of  35  Ibs.  uses  20  Ibs.  of  steam  per  I.H.P.  per  hour. 
Steam  pressure,  100  Ibs.;  feed  temperature,  200°;  coal  costs  $2.50  a  ton  and 
contains  13,500  B.T.U.  per  Ib.  Efficiency  of  the  boiler  plant,  70  per  cent. 
A  gas  engine  is  being  considered  for  the  place.  The  engine  is  8^"  X  14", 
single  acting,  running  223  r.p.m.  and  making  75  explosions  per  minute.  It 
uses  2 1  Ibs.  coal  per  I.H.P.  per  hour.  The  area  of  the  average  indicator  card 
is  1.04  sq.  in.  and  the  length  3.33  in.  Scale  of  spring,  240  Ibs.  The  engines 
are  to  run  ten  hours  a  day,  three  hundred  days  in  the  year.  Gas  producer 
uses  the  same  coal  as  the  boiler  plant.  Which  would  be  the  cheaper  to  run 
and  how  much  per  year?  If  a  Prony  brake  is  placed  on  each  engine,  that  on 
the  steam  engine  having  a  length  of  4  ft.  and  carrying  a  net  weight  of  50  Ibs., 


308  HEAT  ENGINES 

and  that  on  the  gas  engine  having  a  length  of  .63  in.  and  carrying  a  gross 
weight  of  58  Ibs.,  the  tare  being  19  Ibs.,  which  engine  will  develop  the  larger 
output  and  how  much?  Which  has  the  greater  mechanical  efficiency  and 
how  much? 

6.  A  steam  engine  uses  20  Ibs.  of  steam  per  I.H.P.  per  hour  and  develops 
200  H.P.  A  20"  X  24"  single  acting  gas  engine  running  220  r.p.m.  is  being 
considered  for  the  place.  It  uses  10,000  B.T.U.  per  I.H.P.  per  hour  when 
making  105  explosions  per  minute  and  developing  an  average  M.E.P.  of 
100  Ibs.  Efficiency  of  the  boiler  plant,  70  per  cent.;  efficiency  of  gas  pro- 
ducer, 80  per  cent.  Steam  engine  plant  costs  $20,000.  Gas  engine  and  gas 
producer  plant  costs  $30,000.  Cost  of  labor  is  the  same  for  both  plants. 
Coal  costs  $3  a  ton  and  contains  13,000  B.T.U.  perlb.  The  steam  pressure 
in  the  boiler  plant  is  100  Ibs.,  and  the  temperature  of  the  feed  water,  180°. 
If  the  interest  charges  are  5  per  cent.,  and  the  repairs  and  the  depreciation, 
10  per  cent.,  which  would  be  the  cheaper  plant,  and  how  much,  to  run  ten 
hours  a  day  for  three  hundred  days  a  year? 


CHAPTER  XVIII 
ECONOMY  OF  HEAT  ENGINES 

207.  Relative  Economy  of  Heat  Engines. — Primarily  the 
efficiency,  and  in  most  cases,  the  economy,  of  heat  engines 
depends  upon  the  range  of  temperature  of  the  working  medium 
in  the  engine.  As  has  been  shown,  the  thermal  efficiency  of  an 
engine  theoretically  equals 


r, 

where  T\  is  the  initial  absolute  temperature  of  the  working 
medium  and  Tz  is  its  final  absolute  temperature.  In  practice 
it  is  found  that  the  best  heat  engines  are  able  to  realize  actually 
only  about  60  per  cent,  of  the  theoretical  efficiency. 

An  examination  of  the  range  of  temperatures  in  the  various 
forms  of  heat  engines  will  give  some  clue  to  their  probable 
actual  efficiency.  The  following  table  gives  a  general  idea  of 
the  possible  efficiency  of  some  of  the  more  important  prime 
movers. 

TABLE  XXV.    THERMAL  EFFICIENCIES  OF  PRIME  MOVERS 


Range  of 
tempera- 
ture in 
cylinders 

Theore- 
tical 
efficiency 

Probable 
actual 
efficiency 

Average  non-condensing  steam  engine 

116 

14  5 

8.7 

Average  condensing  steam  engine 

226 

27  8 

16  7 

High-pressure  non-condensing  steam  engine.  .  . 
High-pressure  condensing  steam  engine 

194 
279 

22.4 
32  2 

13.4 
19.3 

High-pressure  steam  engine,  superheated  steam 
Average  condensing  steam  turbine,  saturated 
steam 

381 
381 

39.6 
39  6 

23.8 
23  8 

High-pressure  condensing  steam  turbine,  super- 
heated steam  

429 

43.3 

25.7 

Small  gas  engine  

900 

39.5 

19.5 

Large  gas  engine                 

1300 

47.0 

28.0 

Large  gas  engine  high  compression  

1400 

52.2 

31.6 

Diesel  motor  very  high  compression          .... 

1900 

60.0 

36.0 

This  table  gives  some  idea  of  the  development  and  future 

309 


310  HEAT  ENGINES 

possibilities  of  the  various  prime  movers  considering  them  from 
a  standpoint  of  heat  efficiency.  The  internal  combustion  engine 
is  theoretically  approximately  twice  as  efficient  as  the  steam 
engine. 

208.  Commercial  Economy. — Heat  efficiency,  however,  is  not 
the  only  consideration.  In  actual  operation,  the  important 
thing  is  the  cost  to  produce  a  horse-power  for  a  given  period  of 
time.  A  convenient  unit  of  time  is  one  year. 

This  cost  of  production  involves  a  great  many  considerations. 
In  determining  this  cost  the  following  items  should  be  considered : 

(1)  Interest  on  the  capital  invested; 

(2)  Depreciation  of  machinery  and  building  structures; 

(3)  Insurance  and  taxes; 

(4)  Fuel  cost; 

(5)  Labor  of  attendance; 

(6)  Maintenance  and  repairs; 

(7)  Oil,  waste,  water,  and  other  supplies. 

The  first  three  of  these  items  are  called  the  "fixed  charges," 
and  remain  the  same  no  matter  what  the  load  on  the  plant  may 
be.  The  last  four  items  are  the  ''operating  expense,"  and  vary 
with  the  conditions  of  operation.  The  sum  of  the  fixed  charges 
and  operating  expense  is  the  total  operating  cost. 

In  most  plants  the  cost  of  coal  is  from  25  to  30  per  cent,  of 
the  total  operating  expense.  A  saving  in  the  coal  cost  of  operat- 
ing is  not  always  a  saving  in  the  total  cost  of  operating.  This 
saving  may  involve  so  much  increased  cost  of  installation  that 
the  additional  fixed  charges  on  the  new  capital  invested  will 
more  than  offset  the  saving  in  coal.  This  is  well  illustrated  by 
the  condition  which  exists  in  localities  having  very  cheap  coal. 

A  careful  comparison  of  plant-operating  costs  for  a  condensing 
and  a  non-condensing  plant  often  shows  that  the  cost  of  operating 
the  non-condensing  is  less  than  that  of  the  condensing  plant, 
due  to  the  fact  that  the  increased  cost  of  the  condensing  plant 
adds  more  to  the  interest  and  depreciation  charges  than  is  saved 
on  the  cost  of  coal  used,  which  is  less  than  in  a  non-condensing 
plant. 

The  following  table  gives  the  comparative  itemized  costs 
of  operating  for  a  compound  condensing  engine,  a  gas  engine 
with  gas  producer,  and  a  steam  turbine.  These  are  assumed  to 
be  operating  an  electric  generating  unit. 


ECONOMY  OF  HEAT  ENGINES 


311 


Comparison  of  a  1000  B.H.P.  compound  condensing  engine, 
a  1000  B.H.P.  bituminous  gas  producer  and  gas  engine  plant, 
and  a  1000  B.H.P.  steam  turbine.  Bituminous  coal  assumed  to 
cost  $3  per  ton,  with  lower  heat  value  of  12,000  B.T.U.  per  pound 

TABLE  XXVI.     COMPARATIVE  COSTS  PER  RATED  HORSE  POWER 


Reciprocating 
engine 

Gas  engine 

Steam  turbine 

Installation. 
Engine 

$18.00 

8.00 
3.50 

29.50 

27.00 

18.00 
10.00 

$40.00 
3.50 

43.50 

20.00 

18.00 
10.00 

$15.00 

6.00 
5.00 

26.00 

23.50 

14.00 
7.50 

Piping                             

Condensers  and  pumps  

Engine  plant  
Producer  
Boiler                             

12.00 
10.00 
5.00 

20.00 

10.00 
9.00 
4.50 

Chimney,  breeching  and  pumps 
Stokers               

Boiler,  or  producer,  plant.  .  . 
Generator,     switchboard     and 
connections  

4.25 
5.95 
1.70 

4.58 
6.40 
1.83 

3.55 
4.79 
1.42 

Building 

Total  cost  of  plant  
Operation. 

Interest,  5  per  cent  
Depreciation,  7  per  cent  
Insurance  and  taxes.            .    . 

$84.50 

11.90 
43.05 

$91.50 

12.81 
31.85 

$71.00 

9.76 
36.73 

Fixed  charges                

16.20 
2.75 

12.90 

21.60 
2.13 

13.00 

Coal  per  brake  horse-power  per 
year  .  . 

27.00 
2.55 

13.50 

Repairs  3  per  cent. 

Attendance,     oil,     waste     and 
supplies 

Operating  expense  
Total  cost  of  operation  

$54.95 

$44.66 

$46.49 

The  above  table  assumes  the  plant  to  operate  24  hours  per 
day  and  300  days  per  year,  and  the  average  load  to  be  one-half 
of  the  full  rated  load. 


312  HEAT  ENGINES 

As  the  cost  of  coal  increases,  the  gas  engine  and  gas  producer 
will  make  a  more  favorable  showing.  If  full  load  could  be  carried 
for  the  24  hours,  the  showing  will  be  more  favorable  to  the  recip- 
rocating engine.  With  smaller  units  the  cost  of  operation  is  less 
for  the  gas  engine,  as  small  gas  engines  are  more  economical  than 
small  reciprocating  steam  engines,  or  steam  turbines.  With 
large  gas  engines  the  first  cost  is  high  and  the  upkeep  expensive. 


INDEX 


Adiabatic  expansion,  definition  of,  24 
change  of  temperature  during,  28 

Advance  angle,  188,  190 

Advancing  the  spark,  300 

Air,  composition  of,  70 
pump,  240 

Alcohol,  use  of  in  gas  engines,  286 

Ampere,  9 

Angle  of  advance,  188,  190 

Anthracite  coal,  78 

A.  S.  M.  E.  rule  for  finding, 
horse-power  rating  of  boilers,  98 
quality  of  steam,  53 

Automobile  engines,  rated  horse- 
power of,  291 

Babcock  and  Wilcox  boiler,  91 

Barker's  mill,  244 

Barrus,  52 

Barsanti,  271 

Bearing,  162 

Bituminous  coal,  77 

Blow-off  cock,  105 

Boiler  accessories,  105 

Boiler, 

A.  S.  M.  E.  rule  for  horsepower 
rating  of,  98 

Babcock  and  Wilcox  water- tube, 91 

classification  of,  82 

dry-pipe,  85 

economy,  99 

efficiency,  100 

feed  pump  (see  Feed  pump) 

fire-tube,  82 

heating  surface  of  return  fire-tube, 
99 

Heine  water-tube,  94 

horse-power,  actual,  98 
rated,  97 

internally  fired,  87 

locomotive,  89 

losses,  101 

problems,  133 


Boiler,  Rust  water-tube,  95 

Scotch  marine,  87 

setting,  83,  85 

Stirling  water-tube,  92 

tubes,  diameter  of,  99 

tubular,  82 

tubulous,  82 

use  of  tubular,  89 

vertical,  95 

water- tube,  91 

when    use    fire-tube    and    when 
water-tube,  89 

Wickes  water-tube,  95 
Boyle's  law,  11 
Brake  horse-power  (see  Horse-power, 

brake). 

Branca's,  Giovanni,  turbine,  245 
Brasses,  163 
Breeching,  definition  of,  99 

ratio  of  to  grate  surface,  99 
British  Thermal  Unit  (B.T.U.),  5 
Brunton,  109 
Buckeye  riding  cut-off  valve,  205 

Calorific  power  of  fuel,  66,  76,  77, 

78,  79 
Calorimeter,  48 

Barrus  throttling,  52 

Carpenter  separating,  49 
throttling,  50 

coal,  68 

nipple,  48 

"normal  reading"  of,  48 

Peabody  throttling,  50 

problems,  55 

separating,  48 

throttling,  48,  50 
Capacity  of  pump,  181 
Carburetor,  302 
Carnot  cycle,  description  of,  29 

most  efficient  cycle,  34 

reversibility  of,  33 
Carpenter,  49 


313 


314 


INDEX 


Charles'  law,  11 

Chimneys,  boiler  horse-power  of,  130 

brick,  131 

capacity  of,  129 

dpaft  of,  128 

efficiency  of,  130 

height  of,  131 

materials  used  in,  131 

steel,  self-sustaining,  132 

unlined,  132 
Clausius,  10 
Clearance,  149 

per  cent,  of,  149 
Clerk,  Dougal,  269,  278 
Coal  analysis,  64 
proximate,  64 
ultimate,  64 
Coal,  anthracite,  78 

bituminous,  77 

calorimeter,  68  • 

dry,  64 

semi-bituminous,  78 
Cock,  blow-off,  105 

gage,  106 

three-way,  reversing  by  means  of, 
212 

tri-,  107 
Combustion,  air  required  for,  70 

heat  of,  66 

problems,  79 

rate  of,  98 

theoretical  temperature  of,  74 
Compound  engines,   combined  dia- 
gram from,  232 

cross,  227 

cut-off    in    low-pressure    cylinder 
of,  229 

"fore  and  aft,"  228 

horse-power  of,  229 

number  of  cylinders  in,  225 

problems,  234 

ratio  of  cylinders  in,  228 

tandem,  226 
Compound    expansion,   effect   upon 

initial  condensation,  148 
Compounding,  gains  due  to,  225 

losses  due  to.  225 

principal  object  of,  225 


Compression,  149 

Condensation,     initial     (see    Initial 

condensation). 
Condensers,  barometric,  238 

cooling  surface  required  in,  242 

cooling  water  used  by,  240 

for  steam  turbine  use,  242 

increase  in  power  due  to,  242 

jet,  237 

location  of  hot-well  for  use  with, 
238 

surface,  240 

types  of,  236 
Conduction,  7 

Connecting  rod,  effect  of  on  Zeuner 
diagram,  196 

solid-ended,  161 

strap-ended,  161 
Convection,  8 
Cover  plate,  201 
Corliss  engine  card,  174 

trip  gear,  206 

valve,  205 

effect  of  using,  208 
Counter-balance,  162 
Crank-shaft,  162 
Crosby  indicator,  169 
Cross-compound  engine,  227 
Cross-head,  160,  161 

pin,  161 

Curtis  turbine,  255 
"Cushion  steam,"  150 
Cut-off  valve  (see  Valve,  cut-off). 
Cycle,  Beau  de  Rochas,  271 

Carnot,  29 

four-,  272 

Otto,  271 

two-,  272 
"Cylinder  feed,"  150 

Dash-pot,  206 

Davis,  5,  40,  42 

Dead  center,     method     of     placing 

engine  on,  212 
De  Laval  turbine,  251 
Diagram  factor,  143 
Diesel,  269 

motor,  273,  296 


INDEX 


315 


Draft,  chimney,  128 

forced,  132 

induced,  133 

mechanical,  132 

systems  of,  132 
"Dutch  Oven,"  95,  112 
Duty,  180 

Eccentric  rod,  163 

sheave,  163 

strap,  163 

throw  of,  190 
Eccentricity,  188,  190 
Economizers,  cost  of,  126 

description  of,  125 

size  of,  127 

Economy,  engine,  commercial,  310 
relative,  309 

governor,  relative,  216 
Efficiency,  boiler,  100 

boiler  and  grates  combined,  100 

chimney,  130 

fuel,  79 

gas  engine,  275 

producer,  285 

heat  engine,  10,  29,  32 
actual,  180 

mechanical,  179 

turbine,  best,  249 
Energy,  8 

change  in  internal,  due  to  change 

in  temperature,  15 
Engine,  automobile  (see  Automobile 
engine). 

commercial  economy  of,  310 

Corliss  (see  Corliss  engine). 

gas  (see  Gas  engine). 

heat  (see  Heat  engine). 

steam  (see  Steam  engine). 
Equivalent  evaporation,  100 
Evaporation,  equivalent,  100 

factor  of,  100 

per  pound  of  coal,  100 
Exhaust,  heat  lost  in,  145 

lap,  189 

'effect  of  on  Zeuner  diagram,  195 
Expansion,  adiabatic,  24 

change  of  temperature  during,  27 


Expansion,  compound,  effect  of  upon 

initial  condensation,  148 
general  case,  17 
isothermal,  25 
ratio  of,  26,  147 
work  of,  18 

Factor,  diagram,  143 

of  evaporation,  100 
Feed  pump,  location  of,  122 

use  of,  119 

Worthington  boiler,  120 
Feed-water  heaters,  advantages  of, 
123 

closed,  123 

cost  of,  125 

location  of,  125 

open,  123 

types  of,  123 

use  of,  123 
"Fixed  charges,"  310 
Flame  propagation,  287 
Flue  gas,  analysis  of,  71 
Fly-wheel,  223 

Forces  of  impulse  and  reaction,  245 
Frame,  164 

"Free-piston"  gas  engine,  270 
"Fore  and  aft"  compound  engine, 

228 

Fuel,  air  required  for  combustion  of, 
70 

classification  of,  75 

composition  of,  64 

efficiency  of,  79 

gas  engine,  283 

heating  value  of  gas  and  oil,  287 
value  of,  theoretical,  68 

mixtures,  proper,  287 
Fusible  plug,  108 

Gage  cocks,  106 

glass,  105 
Gas  engine, 

Barsanti  and  Matteucci's,  271 

classification,  270 

construction,  details  of,  294 

Diesel's,  273,  280,  296 

efficiency  of,  275 


316 


INDEX 


Gas  engine,  four-cycle,  272 

"free-piston,"  270 

fuels,  283 

governors,  types  of,  300 

history  of  the,  269 

horizontal  vs.  vertical,  303 

horse-power,  actual,  292 
rated,  288 

ignition,  kinds  of,  295 

Langen's,  Otto  and,  270 

Lenoir's,  270 

losses  in  a,  282 

Otto  (and  Langen's),  270 

problems,  307 

two-cycle,  272 

use  of  alcohol  in,  286 

vertical  vs.  horizontal,  303 
Gas  producers,  283 

efficiency  of,  285 

pressure,  284 

suction,  284 

Gas  pump,  Humphrey,  305 
Gear,  Corliss  trip.  206 

Joy,  212 

radial,  211 

reversing,  209,  210 

Walschaert,  211 
Giffard,  M.,  121 
Governor,  automatic,  216 

centrifugal,  220 

design,  relation  of  items  in,  219 

fly-ball,  218 

"hit  and  miss,"  301 

inertia,  221 

isochronous,  222 

mechanism,  217 

quality,  301 

quantity,  301 

shaft,  218,  220 

throttling,  216 

used    with    double-ported    valve, 
203 

variable  cut-off,  216 
Governors,  gas-engine,  types  of,  300 

practical    considerations    in    con- 
nection with,  223 

relative  economy  of,  216 

variation  in  speed  allowable,  223 


Grate  surface  in  stokers,  118 

ratio  of,  to  breeching,  99 

to  heating  surface,  98 
Gutermuth,  148 
Giildner,  269 

Heat,  absorption  of,  13 

added  at  constant  pressure,  22 
at  constant  volume,  22 
general  case,  20 
balance  in  boiler  plant,  101 
capacity,  6 

lost  in  exhaust,  145 
up  stack,  101 
of  fusion  of  ice,  latent,  57 
of  liquid,  38,  40 
of  steam,  latent,  38,  41 

total,  41 
of  superheat,  38 
relation  between,  and  work,  10 
between  specific,  of  constant  pres- 
sure and  of  constant  volume,  15 
specific  (see  Specific  heat), 
theory  of,  1 
unit  of,  5 

Heat  engines,  efficiency  of  (see  Effi- 
ciency, heat  engine), 
ideal,  29 

relative  economy  of,  309 
Heater,  feed-water  (see  Feed-water 

heaters). 

Heating  surface,  definition  of,  98 
of  fire-tube  boilers,  rule  for  finding, 

99 
ratio  of,  to  grate  surface,  98 

to  rated  boiler  horse-power,  99 
Heating  value  of  fuel,  66,  287 
higher,  66 
lower,  66 

of  combustible,  101 
Heine  boiler,  94 
Hero's  turbine,  244 
"Hit  and  miss"  governor,  301 
Hollis,  128 
Horse-power,  boiler, 
actual,  98 

heating  surface  per,  99 
rated,  97 


INDEX 


317 


Horse-power,  brake,  gas  engine,  293 
steam  engine,  178 
chimney,  130 
comparison  of  rated,  for  steam  and 

gas  engines,  292 
engine,  automobile,  rated,  291 
compound,  229 
definition  of,  9 
gas,  actual,  292 

rated,  288 
steam,  indicated,  143,  173 

theoretical,  142 
friction,  179 
of  gas  and   steam   engines,   how 

rated,  292 

Hornsby-Akroyd  oil  engine,  286,  297 
Hot-well,  238 
Humphrey  gas  pump,  305 
"Hunting,"  222 

Ice,  latent  heat  of  fusion  of,  57 
Ignition,  flame,  295 

hot  tube,  295 

jump-spark,  299 

kinds  of  gas  engine,  295 

"make  and  break,"  298 
Impulse,  245 

Indicator,  accuracy  of,  170 
Indicator  cards,  174 

combined,  232 

method  of  taking,  172 

various  forms  of,  213 
Indicator, 

Crosby,  169 

difference  between  steam  and  gas 
engine,  292 

external  spring,  170 

manner  of  connecting  to  engine, 
171 

reducing  motions,  171 

setting  valve  by,  213 

things  determined  by,  168 

Thompson,  170 

use  of,  170 
Initial  condensation,  action  of,  145 

amount  of,  146 

factors  affecting,  146 

graphical  determination  of,  174 


Injectors,  description  of,  121 

location  of,  122 
Isochronous  governor,  222 
Isothermal  expansion,  25 

Joule,  10,  14 
Joy  gear,  212 
Juckes,  John,  110 
Jump-spark  ignition,  299 

Kerr  turbine,  259 
Knock-off  cam,  206 

Langen,  270 
Lap,  exhaust,  189 

effect  of  on  Zeuner  diagram,  195 
Lap  steam,  189 

effect  of  on  Zeuner  diagram,  192 
Lead,  150,  188,  190 
Lenoir,  270 
Lignite,  77 
Losses,  boiler,  101 

due  to  compounding,  225 

gas-engine,  282 

steam-engine,  144 
Lubricators,  165 

Magneto,  297 

Mahler  bomb  coal  calorimeter,  68 
"Make  and  break"  igniter,  298 
Marks,  40,  42 
Matteucci,  271 

Mean  effective  pressure,  143,  173 
Mechanical  draft,  132 
systems  of,  132 

efficiency,  179 

equivalent  of  heat,  10 

mixtures,  56 

stokers  (see  Stokers,  mechanical). 
Meyer  riding  cut-off  valve,  204 
Mines,  U.  S.  Bureau  of,  64,  65 
Mixture  problems,  61 
Mixtures,  fuel,  287 

mechanical,  56 
Moisture  in  steam,  48 
Murphy,  Thomas,  110 

Napier's  rule,  48 


318 


INDEX 


Nozzle,  turbine,  248 

"Normal  reading"  of  calorimeter,  53 


Oil,  vaporization  of,  286 
engine,     Hornsby-Ackroyd, 

297 

for  ships,  205 
"Operating  expense,"  310 
Orsat  apparatus,  72 
Otto,  269 


Parr  coal  calorimeter,  68 

Parsons  turbine,  single-flow,  262 

Peabody,  40,  42,  50 

Peat,  76 

Peclet,  7,  8 

Perfect  gases,  definition  of,  12 

equation  of,  12 

laws  of,  11 

problems  in,  34 
Piston,  160,  161 

position  of  relative  to  valve,  190 

rod,  160,  161 

valve,  198 
Planimeter,  173 
Port  opening,  193 
Power,  9 
Pressure,  absolute,  12 

gage,  12 

gas  producer,  284 

heat  added  at  constant,  22 

mean  effective,  143,  173 

relation    between    volume,    tem- 
perature, and,  12,  27 

specific  heat  of  constant,  7 
Problems,  boiler,  133 

calorimeter,  55 

combustion,  79 

economic,  242 

engine,  gas,  307 
steam,  actual,  184 
compound,  234 
theoretical,  150 

mixture,  61 

perfect  gas,  34 

Producers,  gas  (see  Gas  producers). 
Pumps,  capacity  of,  181 

circulating,  240 


Pumps,  dry  air,  240 

feed  (see  Feed  pumps). 
Pyrometers,  3 


286,      Quality  of  steam,  48,  53 


Radial  gears,  211 

Radiation,  7 

Rankine,  10 

Rateau  turbine,  257 

Ratio  of  expansion,  147,  228 

Re-action,  245 

Reducing  motion,  indicator,  171 

Re-evaporation,  145 

Regnault,  39 

Reversing  gear,  209,  210 

by  means  of  three-way  cock,  212 
Rochas,  271 

Rotation,  changing  direction  of,  209 
Rowland,  10 
"Run  over,"  209 
"Run  under,"  209 
Rust  boiler,  95 

Safety  valve,  size  of,  107 
Sampling  nozzle,  48,  53 
Saturation  curve,  175 
Semi-bituminous  coal,  78 
Smoke,  71 

Southwark-Rateau  turbine,  259 
Specific  heat,  definition  of,  6 
of  constant  pressure,  7 

and  of  constant  volume,  rela- 
tion between,  15 
of  superheated  steam,  40 
of  constant  volume,  7 
theory  of,  5 
Stanley,  109 

Steam,  action  of,  in  turbine,  247 
A.  S.  M.  E.  rules  for  finding  qual- 
ity of,  53 

boiling  point  of,  38 
consumption,    determination    of, 

177 
variation  of  at  different  loads, 

181 

cushion,  150 
dry  saturated,  39 


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